Engine control device

ABSTRACT

A control apparatus for an engine includes an engine, a state quantity setting device, a spark plug, and a controller. The spark plug ignites air-fuel mixture at predetermined ignition timing so that unburned air-fuel mixture combusts by autoignition after start of combustion of the air-fuel mixture by the ignition, and the controller adjusts a heat amount ratio in accordance with an operation state of the engine through change of the ignition timing, the heat amount ratio representing an index associated with a ratio of an amount of heat generated when the air-fuel mixture combusts by flame propagation with respect to a total amount of heat generated when the air-fuel mixture combusts in the combustion chamber.

TECHNICAL FIELD

The present disclosure relates to a control apparatus for an engine.

BACKGROUND ART

Patent Document 1 discloses a compression autoignition type engine inwhich, by compression operation of a piston, air-fuel mixture in acombustion chamber is caused to undergo autoignition so as to becombusted. In this engine, auxiliary ignition is performed by a sparkplug, whereby the combustion speed of air-fuel mixture is controlled. Inthis engine, as the load on the engine decreases, as the number ofrevolutions of the engine increases, or as the oil water temperaturedecreases, air-fuel mixture becomes less likely to undergo autoignition.Therefore, ignition timing by the spark plug is made earlier, wherebyautoignition of the air-fuel mixture is promoted so that time ofcombustion by autoignition is advanced.

Patent Document 2 discloses an engine which combusts air-fuel mixture ina combustion chamber by autoignition in a partial load region. Thisengine promotes autoignition of air-fuel mixture by leaving hot burnedgas in the combustion chamber in an operation region on a low load sidein the partial load region. Furthermore, in the engine, in an operationregion on a high load side in the partial load region, cooled burned gasis introduced into the combustion chamber such that autoignition is lesslikely to occur, and a spark plug performs ignition immediately beforethe compression top dead center.

CITATION LIST Patent Document

PATENT DOCUMENT 1: Japanese Patent No. 3951515

PATENT DOCUMENT 2: Japanese Patent No. 4082292

SUMMARY OF THE INVENTION Technical Problem

In the techniques disclosed in Patent Documents 1 and 2, even if thespark plug performs ignition, it is impossible to perform such controlas to start compression autoignition at desired timing and thus there isa room for improvement in thermal efficiency.

The present disclosure allows unburned air-fuel mixture to beappropriately combusted by autoignition after air-fuel mixture isignited.

Solution to the Problem

The inventors of the present invention have found a combustion mode inwhich SI (Spark Ignition) combustion and CI (Compression Ignition)combustion (or self-ignition (Auto Ignition) combustion) are combined.The SI combustion causes flame propagation that starts by forciblyigniting air-fuel mixture in a combustion chamber. The CI combustionstarts by compression autoignition of air-fuel mixture in a combustionchamber. In the combustion mode in which the SI combustion and the CIcombustion are combined, a spark plug forcibly ignites air-fuel mixturein a combustion chamber, to combust the air-fuel mixture by flamepropagation, and heat generation by the SI combustion enhances thetemperature in the combustion chamber, to combust unburned air-fuelmixture by autoignition. In combustion by flame propagation, pressurevariation is relatively small, and therefore combustion noise can besuppressed. In addition, performing the CI combustion shortens thecombustion period and thus has an advantage in fuel economy improvement,as compared to combustion by flame propagation. The combustion mode inwhich the SI combustion and the CI combustion are combined can improvefuel economy while suppressing combustion noise. In this combustionmode, the CI combustion is controlled by the SI combustion, andtherefore, hereinafter, the combustion mode is referred to as SPCCI(SPark Controlled Compression Ignition) combustion.

However, in combustion by autoignition, if the temperature (hereinafter,referred to as “initial temperature”) in the combustion chamber beforestart of compression varies, autoignition timing greatly changes. If theautoignition timing becomes earlier, start of CI combustion is madeearlier, whereby SI combustion is shortened accordingly. As a result,combustion by flame propagation is not sufficiently performed in thecombustion chamber, and rising of CI combustion becomes steep. Thismight pose an obstacle to suppression of occurrence of combustion noise.On the other hand, if the autoignition timing is delayed, stability ofCI combustion might be lost, leading to increase in unburned fuel orreduction in exhaust gas performance.

The inventors of the present invention have considered controlling theautoignition timing through adjustment of the amount of heat generationby SI combustion. Then, the inventors of the present invention havefound that, if the amount of heat generation by SI combustion isadjusted through change of the ignition timing, the autoignition timingcan be accurately controlled, and thus have reached completion of thetechnique disclosed herein.

Specifically, a technique disclosed here relates to a control apparatusfor an engine. The control apparatus for an engine includes: an enginehaving a combustion chamber; an injector mounted to the engine andconfigured to inject fuel; a spark plug disposed so as to face an insideof the combustion chamber; and a controller connected to the injectorand the spark plug and configured to output control signals to theinjector and the spark plug.

The spark plug ignites air-fuel mixture at predetermined ignition timingso that unburned air-fuel mixture combusts by autoignition after startof combustion of the air-fuel mixture by the ignition, and thecontroller adjusts a heat amount ratio in accordance with an operationstate of the engine through change of the ignition timing, the heatamount ratio representing an index associated with a ratio of an amountof heat generated when the air-fuel mixture combusts by flamepropagation with respect to a total amount of heat generated when theair-fuel mixture combusts in the combustion chamber.

The “combustion chamber” described herein is not limited to a spaceformed when a piston reaches the compression top dead center. The term“combustion chamber” is used to encompass a broader meaning.

In this configuration, the spark plug forcibly ignites air-fuel mixturein the combustion chamber in response to a control signal from thecontroller. The air-fuel mixture combusts by flame propagation, andthereafter, the unburned air-fuel mixture in the combustion chambercombusts by autoignition, whereby combustion is completed. In suchcombustion mode, i.e., SPCCI combustion, the heat amount ratio is lessthan 100%.

If the heat amount ratio is increased in SPCCI combustion, the ratio ofcombustion by flame propagation increases, and thus combustion noise isadvantageously suppressed. On the other hand, if the heat amount ratiois decreased in SPCCI combustion, the ratio of CI combustion increases,and thus fuel economy is advantageously improved. Therefore, if the heatamount ratio is adjusted in accordance with the operation state of theengine, it is possible to improve fuel economy while suppressingoccurrence of combustion noise.

Adjustment of the heat amount ratio is performed through change of thetiming of ignition by the spark plug. If the ignition timing is changed,the start timing of SI combustion changes accordingly, whereby the heatamount ratio can be changed. If the ignition timing is advanced, thestart timing of SI combustion becomes earlier, so that the amount ofheat generation by SI combustion increases and the heat amount ratioincreases. On the other hand, if the ignition timing is retarded, thestart timing of SI combustion is delayed, so that the amount of heatgeneration by SI combustion decreases and the heat amount ratiodecreases.

Then, if the heat amount ratio is changed, the amount of heat generationby SI combustion changes, whereby it is possible to change timing atwhich, through heat generation by flame propagation, the inside of thecombustion chamber reaches a high-temperature environment that inducesautoignition. Thus, it is possible to accurately control timing ofautoignition in SPCCI combustion.

Thus, with the above configuration, it becomes possible to start the CIcombustion at desired timing, whereby unburned air-fuel mixture can beappropriately combusted by autoignition.

The controller may output a control signal to the spark plug so that, asa temperature in the combustion chamber before start of compressiondecreases, the ignition timing is advanced, thereby increasing the heatamount ratio.

In SPCCI combustion, as described above, if the initial temperature inthe combustion chamber varies, timing of autoignition greatly changes.If the initial temperature in the combustion chamber is high, timing ofautoignition becomes earlier, so that SI combustion is shortened. As aresult, rising of CI combustion becomes steep, so that combustion noiseis likely to increase. If the initial temperature in the combustionchamber is low, timing of autoignition is delayed and stability of CIcombustion is lost. As a result, increase in unburned fuel or reductionof exhaust gas performance can occur.

In this regard, if the heat amount ratio is adjusted in accordance withthe initial temperature in the combustion chamber as described above,variation in the initial temperature in the combustion chamber iscompensated for by heat generation by flame propagation, and thus,autoignition of unburned air-fuel mixture can be caused at desiredtiming.

Another technique disclosed here relates to a control apparatus for anengine. The control apparatus for an engine includes: an engine having acombustion chamber; a state quantity setting device provided to theengine and configured to adjust introduction of fresh air and burned gasinto the combustion chamber; an injector mounted to the engine andconfigured to perform injection; a spark plug disposed so as to face aninside of the combustion chamber; and a controller connected to thestate quantity setting device, the injector, and the spark plug, andconfigured to output control signals to the state quantity settingdevice, the injector, and the spark plug.

The spark plug ignites air-fuel mixture at predetermined ignition timingso that unburned air-fuel mixture combusts by autoignition after startof combustion of the air-fuel mixture by the ignition. The statequantity setting device adjusts operation quantity relevant to atemperature in the combustion chamber before start of compression, inresponse to a control signal from the controller. The controller adjustsa heat amount ratio in accordance with an operation state of the enginethrough change of the operation quantity by the state quantity settingdevice and change of the ignition timing, the heat amount ratiorepresenting an index associated with a ratio of an amount of heatgenerated when the air-fuel mixture combusts by flame propagation withrespect to a total amount of heat generated when the air-fuel mixturecombusts in the combustion chamber.

In this configuration, when the operation state of the engine is in apredetermined operation region, the spark plug forcibly ignites air-fuelmixture in the combustion chamber. The air-fuel mixture combusts byflame propagation, and thereafter, unburned air-fuel mixture in thecombustion chamber combusts by autoignition, whereby combustion iscompleted. In such combustion mode, i.e., SPCCI combustion, the heatamount ratio is less than 100%.

If the heat amount ratio is increased in SPCCI combustion, the ratio ofcombustion by flame propagation increases, and thus combustion noise isadvantageously suppressed. On the other hand, if the heat amount ratiois decreased in SPCCI combustion, the ratio of CI combustion increases,and thus fuel economy is advantageously improved. Therefore, byadjusting the heat amount ratio in accordance with the operation stateof the engine, it is possible to improve fuel economy while suppressingoccurrence of combustion noise.

Adjustment of the heat amount ratio is performed through change of theoperation quantity relevant to the initial temperature in the combustionchamber by the state quantity setting device, and change of timing ofignition by the spark plug.

If the operation quantity relevant to the initial temperature in thecombustion chamber is changed by the state quantity setting device, thetemperature difference until, through heat generation by SI combustion,the inside of the combustion chamber reaches a high-temperatureenvironment that induces autoignition, varies along with change of theinitial temperature in the combustion chamber, and accordingly, thestart timing of CI combustion changes. That is, timing of shifting fromcombustion by only SI combustion to combustion including CI combustionis changed. Thus, the heat amount ratio can be changed.

If the initial temperature in the combustion chamber becomes high, thetemperature difference until, through heat generation by SI combustion,the inside of the combustion chamber reaches a high-temperatureenvironment that induces autoignition, decreases. Therefore, a perioduntil autoignition is reached from the start of SI combustion byignition by the spark plug, is shortened. As a result, timing ofshifting from combustion by only SI combustion to combustion includingCI combustion becomes earlier, so that the heat amount ratio decreases.

On the other hand, if the initial temperature in the combustion chamberbecomes low, the temperature difference until, through heat generationby SI combustion, the inside of the combustion chamber reaches ahigh-temperature environment that induces autoignition, increases.Therefore, a period until autoignition is reached from the start of SIcombustion by ignition by the spark plug, is prolonged. As a result,timing of shifting from combustion by only SI combustion to combustionincluding CI combustion is delayed, so that the heat amount ratioincreases.

If the ignition timing is changed, the start timing of SI combustionchanges accordingly, whereby the heat amount ratio can be changed. Ifthe ignition timing is advanced, start of SI combustion becomes earlier,so that the amount of heat generation by SI combustion increases and theheat amount ratio increases. On the other hand, if the ignition timingis retarded, start of SI combustion is delayed, so that the amount ofheat generation by SI combustion decreases and the heat amount ratiodecreases.

If the heat amount ratio is changed through change of the operationquantity relevant to the initial temperature in the combustion chamberby the state quantity setting device and change of timing of ignition bythe spark plug as described above, the amount of heat generation by SIcombustion changes, whereby it is possible to change timing at which,through heat generation by flame propagation, the inside of thecombustion chamber reaches a high-temperature environment that inducesautoignition of unburned air-fuel mixture. Thus, it is possible toaccurately control timing of autoignition in SPCCI combustion.

Here, change of the initial temperature in the combustion chamber has agreater influence on the autoignition timing of unburned air-fuelmixture than change of timing of ignition by the spark plug. In view ofthis, it is conceivable that adjustment of the heat amount ratio isperformed such that, for example, the heat amount ratio is roughlyadjusted through adjustment of the operation quantity relevant to theinitial temperature in the combustion chamber by the state quantitysetting device, and the heat amount ratio is finely adjusted throughchange of timing of ignition by the spark plug as described above. Inthis way, by performing adjustment of the heat amount ratio in twostages, it is possible to accurately control the autoignition timing. Asa result, the engine can accurately achieve target SPCCI combustioncorresponding to the operation state.

Thus, with the above configuration, it becomes possible to start the CIcombustion at desired timing, whereby unburned air-fuel mixture can beappropriately combusted by autoignition.

When load on the engine is high, the controller may output a controlsignal for adjusting the operation quantity to the state quantitysetting device so that the temperature in the combustion chamber beforestart of compression becomes lower than when the load is low, therebyincreasing the heat amount ratio.

In SPCCI combustion, the initial temperature in the combustion chambertends to increase as the load on the engine increases. If the initialtemperature in the combustion chamber increases, as described above,timing of shifting from combustion by only SI combustion to combustionincluding CI combustion becomes earlier, so that the heat amount ratiodecreases. If the heat amount ratio decreases, rising of CI combustionbecomes steep as the ratio of CI combustion increases. Thus, combustionnoise increases.

In this regard, when the load on the engine is high, the initialtemperature in the combustion chamber is made lower than when the loadis low, through adjustment of the operation quantity relevant to theinitial temperature in the combustion chamber by the state quantitysetting device. Thus, timing of shifting from combustion by only SIcombustion to combustion including CI combustion is delayed, whereby theheat amount ratio can be increased. Thus, combustion by flamepropagation is ensured, and CI combustion can be prevented from becomingexcessively steep. As a result, it is possible to suitably suppressoccurrence of combustion noise while achieving fuel economy improvementby CI combustion.

When the load on the engine is high, the controller may output a controlsignal to the spark plug so that the ignition timing is advanced ascompared to when the load is low, thereby increasing the heat amountratio.

If the ignition timing by the spark plug is advanced in addition todecrease of the initial temperature in the combustion chamber asdescribed above, start of SI combustion is made earlier while start ofCI combustion is delayed. Therefore, when the load on the engine ishigh, a period during which SI combustion is performed is made longerthan when the load is low, whereby the heat amount ratio in SPCCIcombustion can be further increased. Thus, it is possible tosufficiently suppress occurrence of combustion noise when the load onthe engine is high.

The state quantity setting device may have a supercharging systemprovided to the engine and configured to perform supercharging with gasto be introduced into the combustion chamber, and the superchargingsystem may be configured such that, in response to a control signal fromthe controller, the supercharging system performs supercharging when theload on the engine is high, and does not perform supercharging when theload is low.

For the supercharging system, for example, a mechanical superchargerdriven by the engine may be employed. The mechanical supercharger canswitch between supercharging and non-supercharging.

When the load on the engine is relatively low, the amount of fuel issmall, and therefore it is possible to introduce, into the combustionchamber, fresh air needed for bringing the excess air ratio λ ofair-fuel mixture into a desired state, without performing supercharging.In the case of employing a mechanical supercharger for the superchargingsystem, fuel economy is improved by performing no supercharging.

If the load on the engine becomes high, the amount of fuel increases.Therefore, fresh air needed for bringing the excess air ratio λ into adesired state increases, and it is necessary to introduce a large amountof burned gas into the combustion chamber in order to bring the G/F ofair-fuel mixture into a desired state. When the load on the engine isrelatively high, it is possible to bring the inside of the combustionchamber into a desired state regarding the excess air ratio λ and theG/F, by performing supercharging.

In a case of not performing supercharging using the superchargingsystem, the controller may output a control signal for adjusting theoperation quantity to the state quantity setting device so that thetemperature in the combustion chamber before start of compressiondecreases as the load on the engine increases, thereby increasing theheat amount ratio.

As described above, in SPCCI combustion, the initial temperature in thecombustion chamber tends to increase as the load on the engineincreases. If the initial temperature in the combustion chamberincreases, the heat amount ratio decreases, resulting in increase incombustion noise.

In this regard, if the initial temperature in the combustion chamber isdecreased through adjustment of the operation quantity by the statequantity setting device as the load on the engine increases, shiftingfrom combustion by only SI combustion to combustion including CIcombustion is delayed, whereby the amount of heat generation by SIcombustion can be increased. In this way, it is possible to increase theheat amount ratio in SPCCI combustion as the load on the engineincreases. As a result, in the case of not performing supercharging, itis possible to suitably suppress occurrence of combustion noise.

In a case of not performing supercharging using the superchargingsystem, the controller may output a control signal to the spark plug sothat the ignition timing is advanced as the load on the engineincreases, thereby increasing the heat amount ratio.

If the ignition timing by the spark plug is advanced in addition todecrease of the initial temperature in the combustion chamber asdescribed above, start of SI combustion is made earlier while start ofCI combustion is delayed. Therefore, as the load on the engineincreases, a period during which SI combustion is performed isprolonged, whereby the heat amount ratio in SPCCI combustion can befurther increased. Thus, in the case of not performing supercharging,occurrence of combustion noise can be suppressed as much as possible.

In the case of performing supercharging using the supercharging system,the controller may output a control signal to the spark plug so that theignition timing is advanced as the load on the engine increases, therebymaking the heat amount ratio constant with respect to change in the loadon the engine.

In the case of performing supercharging, if the heat amount ratio ismade constant with respect to change in the load on the engine by theignition timing being advanced, the amount of heat generation by SIcombustion and the amount of heat generation by CI combustion are bothincreased as the load on the engine increases. Therefore, combustionnoise tends to increase by an amount corresponding to increase in thepeak of CI combustion. However, in the case of performing supercharging,the load is relatively high as compared to the case of not performingsupercharging, and accordingly, a certain level of combustion noise ispermissible. Therefore, by ensuring the amount of heat generation by CIcombustion, the combustion period can be shortened, whereby a fueleconomy improving effect can be suitably obtained.

The controller may output control signals to the state quantity settingdevice and the injector, to set a G/F that represents an indexassociated with a weight ratio between total gas including burned gas inthe combustion chamber, and fuel, such that the G/F is in a range from18 to 50.

Through studies by the inventors of the present invention, it has beenfound that, in SPCCI combustion in which SI combustion and CI combustionare combined, the SPCCI combustion is appropriately performed if the G/Fof air-fuel mixture is set in a range from 18 to 50.

Furthermore, when a state inside the combustion chamber satisfies18≤G/F, a dilution rate of the air-fuel mixture is great, whereby fueleconomy performance of the engine is improved. Furthermore, combustionnoise caused by knocking can be assuredly avoided.

The controller may output control signals to the state quantity settingdevice and the injector, to set an excess air ratio λ of the air-fuelmixture to 1.0±0.2.

By setting the G/F of air-fuel mixture in a range from 18 to 50 andsetting λ to 1.0±0.2, it is possible to accurately control theautoignition timing in SPCCI combustion.

In SPCCI combustion, even if variation occurs in the temperature in thecombustion chamber before start of compression, it is possible to absorbthe variation in the temperature before start of compression byadjusting the amount of heat generation in SI combustion. If timing ofstarting SI combustion is adjusted by, for example, adjustment ofignition timing in accordance with the temperature in the combustionchamber before start of compression, the unburned air-fuel mixture canbe caused to undergo autoignition at target timing.

However, in order to accurately control the autoignition timing in SPCCIcombustion, the autoignition timing has to be changed in response to thechanging of the ignition timing. It is preferable that sensitivity ofchange of the autoignition timing with respect to change of the ignitiontiming is high.

Through studies by the inventors of the present invention, it has beenfound that, if the state inside the combustion chamber is set so that λof the air-fuel mixture is 1.0±0.2 and the G/F of the air-fuel mixtureis in a range from 18 to 50, the SI combustion is stabilized and as aresult, the autoignition timing changes in response to changing of theignition timing. That is, in SPCCI combustion, the autoignition timingcan be accurately controlled.

Further, if λ is set to 1.0±0.2, it becomes possible to purify theexhaust gas by a three-way catalyst mounted to the exhaust passage ofthe engine.

Thus, with the above configuration, it is possible to accurately controlthe autoignition timing in the SPCCI combustion in which SI combustionand CI combustion are combined, while enhancing fuel economy performanceand obtaining good exhaust gas performance.

By controlling the autoignition timing by SI combustion, it is possibleto cause the unburned air-fuel mixture to undergo autoignition at timingoptimum in terms of fuel economy while suppressing combustion noise,even if the temperature in the combustion chamber before start ofcompression varies.

A state inside the combustion chamber at the ignition timing may satisfyat least one of a condition that a temperature is in a range from 570 Kto 800 K, and a condition that a pressure is in a range from 400 kPa to920 kPa.

This can stabilize the SPCCI combustion.

A state inside the combustion chamber at the ignition timing may satisfya condition that a swirl ratio is 4 or greater.

Strengthening swirl flow in the combustion chamber enables the SPCCIcombustion to be performed stably.

A geometrical compression ratio of the engine may be 13 or greater. InSPCCI combustion, since spark ignition is performed, it is not necessaryto, for autoignition of air-fuel mixture, greatly increase thetemperature in the combustion chamber when the piston reaches thecompression top dead center. If the geometrical compression ratio is setto be small, there is an advantage in reduction of cooling loss andreduction of mechanical loss of the engine.

Advantages of the Invention

As described above, in the above control apparatus for the engine,unburned air-fuel mixture appropriately combusts by autoignition afterignition of air-fuel mixture.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 illustrates a configuration of an engine.

FIG. 2 illustrates a configuration of a combustion chamber, and theupper view thereof corresponds to a view of the combustion chamber in aplaner view and the lower view thereof is a cross-sectional view takenalong II-II.

FIG. 3 is a plan view of a configuration of the combustion chamber andan intake system.

FIG. 4 is a block diagram illustrating a configuration of a controlapparatus for the engine.

FIG. 5 illustrates a rig tester for measuring a swirl ratio.

FIG. 6 illustrates a relationship between a swirl ratio and an openingratio of a secondary passage.

FIG. 7 illustrates, in the upper diagram, an operation region map of theengine and illustrates, in the lower diagram, an operation region mapdifferent from that in the upper diagram.

FIG. 8 conceptually illustrates, in the upper diagram, change of a heatgeneration rate in SPCCI combustion in which SI combustion and CIcombustion are combined, illustrates, in the intermediate diagram,definition of an SI rate in SPCCI combustion, and illustrates, in thelower diagram, another definition of an SI rate in SPCCI combustion.

FIG. 9 illustrates change of an SI rate, change of a state quantity inthe combustion chamber, change of an overlap period of an intake valveand an exhaust valve, and changes of fuel injection timing and ignitiontiming, according to whether load on the engine is high or low.

FIG. 10 illustrates, in the upper diagram, change of a combustionwaveform with respect to increase of load on the engine innon-supercharge SPCCI combustion, and illustrates, in the lower diagram,change of a combustion waveform with respect to increase of load on theengine in supercharge SPCCI combustion.

FIG. 11 is a flow chart showing a procedure of control, of the engine,performed by an ECU.

FIG. 12 conceptually illustrates control of adjustment of an SI rate.

FIG. 13 illustrates, in the upper diagram, a relationship between a G/Fof air-fuel mixture, and turbulent energy necessary for obtaining adesired turbulent combustion speed, illustrates, in the intermediatediagram, a relationship between a G/F of air-fuel mixture andtemperature in the combustion chamber for obtaining the necessaryturbulent energy illustrated in the upper diagram, and illustrates, inthe lower diagram, a relationship between a G/F of air-fuel mixture andpressure in the combustion chamber for obtaining the necessary turbulentenergy illustrated in the upper diagram.

FIG. 14 is a contour view illustrating a contour on a plane in which thevertical axis represents an EGR ratio of air-fuel mixture and thehorizontal axis represents an A/F of the air-fuel mixture, illustratinga change rate of change of autoignition timing with respect to change ofignition timing in SPCCI combustion.

FIG. 15 illustrates a method for considering a relationship, between anexternal EGR ratio of an SI part and a total EGR ratio in the entiretyof the combustion chamber, which is necessary for causing SPCCIcombustion in a state where a G/F of air-fuel mixture stratifies in thecombustion chamber.

FIG. 16 illustrates a relationship, between the G/F of the SI part andthe G/F in the entirety of the combustion chamber, which is necessaryfor causing SPCCI combustion in a state where the G/F of air-fuelmixture stratifies in the combustion chamber.

FIG. 17 illustrates a fuel injection time, an ignition time, and acombustion waveform in each operation state in the operation region mapshown in the lower diagram of FIG. 7.

FIG. 18 illustrates difference among combustion waveforms when theignition timing is changed in accordance with the initial temperature inthe combustion chamber.

FIG. 19 illustrates change in the ignition timing with respect to theinitial temperature in the combustion chamber, in a region in which theload is a predetermined load or higher in a non-supercharge SPCCIregion.

FIG. 20 illustrates change in an SI rate with respect to the ignitiontiming, in a region in which the load is a predetermined load or higherin the non-supercharge SPCCI region.

FIG. 21 illustrates change in the ignition timing with respect to theload on the engine.

FIG. 22 illustrates change in the initial temperature in the combustionchamber with respect to the load on the engine, in a region in which theload is a predetermined load or higher in the non-supercharge SPCCIregion.

DESCRIPTION OF EMBODIMENTS

Hereinafter, embodiments of a control apparatus for an engine will bedescribed in detail with reference to the drawings. Described below isan example of the control apparatus for the engine. FIG. 1 illustrates aconfiguration of the engine. FIG. 2 is a cross-sectional view of aconfiguration of a combustion chamber. The upper view of FIG. 2corresponds to a view of the combustion chamber in a planer view, andthe lower view thereof is a cross-sectional view taken along II-II. FIG.3 illustrates a configuration of the combustion chamber and an intakesystem. In FIG. 1, the intake side is the left side on the drawingsheet, and the exhaust side is the right side on the drawing sheet. InFIG. 2 and FIG. 3, the intake side is the right side on the drawingsheet, and the exhaust side is the left side on the drawing sheet. FIG.4 is a block diagram illustrating a configuration of the controlapparatus for the engine.

The engine 1 is a four-stroke engine that operates by repeating anintake stroke, a compression stroke, an expansion stroke, and an exhauststroke in a combustion chamber 17. The engine 1 is mounted to afour-wheeled automobile. By operation of the engine 1, the automobileruns. Fuel for the engine 1 is gasoline in this exemplary configuration.The fuel may be gasoline that contains bioethanol or the like. The fuelfor the engine 1 may be any fuel when the fuel is liquid fuel thatcontains at least gasoline.

(Configuration of Engine)

The engine 1 includes a cylinder block 12 and a cylinder head 13 placedthereon. A plurality of cylinders 11 are formed in the cylinder block12. In FIG. 1 and FIG. 2, one cylinder 11 is merely illustrated. Theengine 1 is a multi-cylinder engine.

In each of the cylinders 11, a piston 3 is slidably inserted. The piston3 is connected to a crankshaft 15 via a connecting rod 14. The piston 3,the cylinder 11, and the cylinder head 13 define the combustion chamber17. The “combustion chamber” is not limited to a space formed when thepiston 3 reaches the compression top dead center. The term “combustionchamber” may be used to encompass a broader meaning. That is, the“combustion chamber” may be a space formed by the piston 3, the cylinder11, and the cylinder head 13 regardless of the position of the piston 3.

The lower surface of the cylinder head 13, that is, the ceiling surfaceof the combustion chamber 17 is formed by an inclined surface 1311 andan inclined surface 1312, as shown in FIG. 2. The inclined surface 1311is inclined upward from the intake side toward an injection axis X2 ofan injector 6 described below. The inclined surface 1312 is inclinedupward from the exhaust side toward the injection axis X2. The ceilingsurface of the combustion chamber 17 has a so-called pent-roof shape.

The upper surface of the piston 3 is raised toward the ceiling surfaceof the combustion chamber 17. A cavity 31 is formed in the upper surfaceof the piston 3. The cavity 31 is recessed from the upper surface of thepiston 3. The cavity 31 is shallow-dish-shaped. The cavity 31 opposesthe injector 6 described below when the piston 3 is positioned at ornear the compression top dead center.

The center of the cavity 31 is shifted to the exhaust side with respectto a central axis X1 of the cylinder 11. The center of the cavity 31 isaligned with the injection axis X2 of the injector 6. The cavity 31 hasa projection 311. The projection 311 is aligned with the injection axisX2 of the injector 6. The projection 311 has an almost conic shape. Theprojection 311 extends upward from the bottom of the cavity 31 toward aceiling surface of the cylinder 11.

The cavity 31 also has a depressed portion 312 formed around theprojection 311. The depressed portion 312 is formed so as to surroundthe entire circumference of the projection 311. The cavity 31 has asymmetric shape about the injection axis X2.

The circumferential side surface of the depressed portion 312 isinclined relative to the injection axis X2 from the bottom surface ofthe cavity 31 toward the opening of the cavity 31. The inner diameter ofthe cavity 31 in the depressed portion 312 is gradually increased fromthe bottom of the cavity 31 toward the opening of the cavity 31.

The shape of the combustion chamber 17 is not limited to the shapeillustrated in FIG. 2. For example, the shape of the cavity 31, theshape of the upper surface of the piston 3, the shape of the ceilingsurface of the combustion chamber 17, and the like can be changed asappropriate.

The engine 1 has a geometrical compression ratio set in a range from 13to 30. As described below, the engine 1 performs, in a part of theoperation region, SPCCI combustion in which SI combustion and CIcombustion are combined. In the SPCCI combustion, the CI combustion iscontrolled by utilizing heat generation and increase in pressure in theSI combustion. In the engine 1, a temperature (that is, compression endtemperature) in the combustion chamber 17 in the case of the piston 3having reached the compression top dead center need not be high forautoignition of air-fuel mixture. That is, while the engine 1 performsthe CI combustion, the geometrical compression ratio can be set so as tobe relatively small. When the geometrical compression ratio is set to besmall, cooling loss and mechanical loss are advantageously reduced. Forexample, the engine 1 may have the geometrical compression ratio of 14to 17 in regular specifications (the octane number of the fuel is about91), and the geometrical compression ratio of 15 to 18 in thehigh-octane specifications (the octane number of the fuel is about 96).

The cylinder head 13 has an intake port 18 formed for each cylinder 11.The intake port 18 has two intake ports that are a first intake port 181and a second intake port 182, as shown in FIG. 3. The first intake port181 and the second intake port 182 are aligned in the axial direction ofthe crankshaft 15, that is, the front-rear direction of the engine 1.The intake port 18 communicates with the combustion chamber 17. Theintake port 18 is a so-called tumble port, which is not shown in detail.That is, the intake port 18 has a shape that allows tumble flow to beformed in the combustion chamber 17.

An intake valve 21 is disposed in the intake port 18. The intake valve21 opens and closes a portion between the combustion chamber 17 and theintake port 18. The intake valve 21 is opened and closed atpredetermined timing by a valve mechanism. The valve mechanism may be avariable valve mechanism that can vary valve timing and/or valve lift.In this exemplary configuration, as shown in FIG. 4, the variable valvemechanism has an intake electric S-VT (Sequential-Valve Timing) 23. Theintake electric S-VT 23 is configured to sequentially change arotational phase of an intake cam shaft in a predetermined angularrange. Thus, the opening time and the closing time of the intake valve21 are sequentially changed. The intake valve mechanism may have an oilhydraulic S-VT instead of the electric S-VT.

The cylinder head 13 also has an exhaust port 19 formed for eachcylinder 11. The exhaust port 19 also has two exhaust ports that are afirst exhaust port 191 and a second exhaust port 192, as shown in FIG.3. The first exhaust port 191 and the second exhaust port 192 arealigned in the front-rear direction of the engine 1. The exhaust port 19communicates with the combustion chamber 17.

An exhaust valve 22 is disposed in the exhaust port 19. The exhaustvalve 22 opens and closes a portion between the combustion chamber 17and the exhaust port 19. The exhaust valve 22 is opened and closed atpredetermined timing by a valve mechanism. The valve mechanism may be avariable valve mechanism that can vary valve timing and/or valve lift.In this exemplary configuration, as shown in FIG. 4, the variable valvemechanism has an exhaust electric S-VT 24. The exhaust electric S-VT 24is configured to sequentially change a rotational phase of an exhaustcam shaft in a predetermined angular range. Thus, the opening time andthe closing time of the exhaust valve 22 are sequentially changed. Theexhaust valve mechanism may have an oil hydraulic S-VT instead of theelectric S-VT.

As described below in detail, in the engine 1, the length of an overlapperiod for an opening time of the intake valve 21 and a closing time ofthe exhaust valve 22, is adjusted by the intake electric S-VT 23 and theexhaust electric S-VT 24. Thus, residual gas in the combustion chamber17 is scavenged. Furthermore, by adjusting the length of the overlapperiod, internal EGR (Exhaust Gas Recirculation) gas is introduced intothe combustion chamber 17, or is confined in the combustion chamber 17.In this exemplary configuration, the intake electric S-VT 23 and theexhaust electric S-VT 24 form an internal EGR system that is onecomponent of a state quantity setting device. The internal EGR systemmay not be formed by the S-VT.

The cylinder head 13 has the injector 6 mounted for each cylinder 11.The injector 6 is configured to inject fuel directly into the combustionchamber 17. The injector 6 is disposed in the valley portion of thepent-roof where the inclined surface 1311 on the intake side and theinclined surface 1312 on the exhaust side intersect each other. As shownin FIG. 2, the injector 6 has the injection axis X2 that is disposedcloser to the exhaust side than the central axis X1 of the cylinder 11is. The injection axis X2 of the injector 6 is parallel to the centralaxis X1. The injection axis X2 of the injector 6 is aligned with theprojection 311 of the cavity 31 as described above. The injector 6opposes the cavity 31. The injection axis X2 of the injector 6 maycoincide with the central axis X1 of the cylinder 11. Also in this case,the injection axis X2 of the injector 6 is preferably aligned with theprojection 311 of the cavity 31.

The injector 6 is implemented by a multi-hole fuel injection valvehaving a plurality of holes, which is not shown in detail. The injector6 injects fuel so as to spread fuel spray radially from the center ofthe combustion chamber 17 as indicated by alternate long and two shortdashes lines in FIG. 2. In this exemplary configuration, the injector 6has ten holes and the holes are disposed at regular angular intervals inthe circumferential direction. The axis of the hole is positioned so asto be shifted, relative to a spark plug 25 described below, in thecircumferential direction, as indicated in the upper view of FIG. 2.That is, the spark plug 25 is disposed between the axes of two holesadjacent to each other. Thus, spray of fuel injected from the injector 6is prevented from being applied directly to the spark plug 25 and makingan electrode wet.

As described below, the injector 6 may inject fuel at timing when thepiston 3 is positioned at or near the compression top dead center. Inthis case, when the injector 6 injects fuel, fuel spray flows downwardalong the projection 311 of the cavity 31 while mixing with fresh air,and flows so as to spread radially from the center of the combustionchamber 17 along the bottom surface and the circumferential side surfaceof the depressed portion 312 in the radially outward direction.Thereafter, air-fuel mixture reaches the opening of the cavity 31, andflows from the radially outer side, along the inclined surface 1311 onthe intake side and the inclined surface 1312 on the exhaust side,toward the center of the combustion chamber 17.

The injector 6 may not be a multi-hole injector. The injector 6 may bean outward-opening valve type injector.

A fuel supply system 61 is connected to the injector 6. The fuel supplysystem 61 has a fuel tank 63 configured to store fuel, and a fuel supplypassage 62 that connects the fuel tank 63 and the injector 6 to eachother. In the fuel supply passage 62, a fuel pump 65 and a common rail64 are disposed. The fuel pump 65 feeds fuel to the common rail 64 underpressure. In this exemplary configuration, the fuel pump 65 is aplunger-type pump driven by the crankshaft 15. The common rail 64 isconfigured to store fuel having been fed under pressure from the fuelpump 65, with a high fuel pressure. When the injector 6 opens, the fuelstored in the common rail 64 is injected through the hole of theinjector 6 into the combustion chamber 17. The fuel supply system 61 isconfigured to be capable of supplying, to the injector 6, fuel under ahigh pressure of 30 MPa or higher. The highest fuel pressure in the fuelsupply system 61 may be, for example, about 120 MPa. The pressure offuel to be supplied to the injector 6 may be changed according to anoperation state of the engine 1. The configuration of the fuel supplysystem 61 is not limited to the above-described one.

The cylinder head 13 has the spark plug 25 mounted for each cylinder 11.The spark plug 25 forcibly ignites air-fuel mixture in the combustionchamber 17. In this exemplary configuration, the spark plug 25 isdisposed closer to the intake side than the central axis X1 of thecylinder 11 is. The spark plug 25 is disposed between the two intakeports 18. The spark plug 25 is mounted to the cylinder head 13 so as tobe inclined from the upper side to the lower side toward the center ofthe combustion chamber 17. The electrode of the spark plug 25 faces theinside of the combustion chamber 17 and is disposed near the ceilingsurface of the combustion chamber 17 as shown in FIG. 2. The positionwhere the spark plug 25 is disposed is not limited to the exemplaryposition shown in FIG. 2. The spark plug 25 may be disposed closer tothe exhaust side than the central axis X1 of the cylinder 11 is.Furthermore, the spark plug 25 may be disposed on the central axis X1 ofthe cylinder 11, and the injector 6 may be disposed closer to the intakeside or the exhaust side than the central axis X1 is.

An intake passage 40 is connected to one side surface of the engine 1.The intake passage 40 communicates with the intake port 18 of eachcylinder 11. Through the intake passage 40, gas to be introduced intothe combustion chamber 17 flows. An air cleaner 41 for filtering freshair is disposed at the upstream end portion of the intake passage 40. Asurge tank 42 is disposed near the downstream end of the intake passage40. The intake passage 40 disposed downstream of the surge tank 42 formsindependent passages that diverge for the respective cylinders 11, whichis not shown in detail. The downstream end of the independent passage isconnected to the intake port 18 of each cylinder 11.

A throttle valve 43 is disposed between the air cleaner 41 and the surgetank 42 in the intake passage 40. A valve opening degree of the throttlevalve 43 is adjusted, whereby an amount of fresh air to be introducedinto the combustion chamber 17 is adjusted. The throttle valve 43 is onecomponent of the state quantity setting device.

A supercharger 44 is disposed downward of the throttle valve 43 in theintake passage 40. The supercharger 44 is configured to performsupercharging with gas that is to be introduced into the combustionchamber 17. In this exemplary configuration, the supercharger 44 is amechanical supercharger driven by the engine 1. The mechanicalsupercharger 44 may be, for example, of a Lysholm type. The mechanicalsupercharger 44 may have any structure. The mechanical supercharger 44may be of a Roots type, a vane type, or a centrifugal type. Thesupercharger may be an electric supercharger, or a turbosuperchargerdriven by exhaust energy.

An electromagnetic clutch 45 is disposed between the supercharger 44 andthe engine 1. Between the supercharger 44 and the engine 1, theelectromagnetic clutch 45 transmits driving force from the engine 1 tothe supercharger 44 and interrupts transmission of driving force. Asdescribed below, an ECU 10 switches between disengagement and engagementof the electromagnetic clutch 45, whereby the supercharger 44 switchesbetween on and off. That is, the engine 1 is configured to switchbetween a state where the supercharger 44 performs supercharging withgas that is to be introduced into the combustion chamber 17 and a statewhere the supercharger 44 does not perform supercharging with gas thatis to be introduced into the combustion chamber 17.

An intercooler 46 is disposed downward of the supercharger 44 in theintake passage 40. The intercooler 46 is configured to cool gascompressed by the supercharger 44. The intercooler 46 may be, forexample, of a water-cooling type.

A bypass passage 47 is connected to the intake passage 40. The bypasspassage 47 connects the upstream portion of the supercharger 44 and thedownstream portion of the intercooler 46 to each other in the intakepassage 40 so as to bypass the supercharger 44 and the intercooler 46.More specifically, the bypass passage 47 is connected to the surge tank42. An air bypass valve 48 is disposed in the bypass passage 47. The airbypass valve 48 adjusts a flow rate of gas that flows in the bypasspassage 47.

When the supercharger 44 is off (that is, the electromagnetic clutch 45is disengaged), the air bypass valve 48 is fully opened. Thus, gas thatflows in the intake passage 40 bypasses the supercharger 44 and isintroduced into the combustion chamber 17 of the engine 1. The engine 1operates in a non-supercharged state, that is, by natural aspiration.

When the supercharger 44 is on (that is, the electromagnetic clutch 45is engaged), a part of gas that passes through the supercharger 44 flowsback to the upstream side of the supercharger through the bypass passage47. By the opening degree of the air bypass valve 48 being adjusted, anamount of backflow can be adjusted, whereby boost pressure for gas to beintroduced into the combustion chamber 17 can be adjusted. A time whensupercharging is performed may be defined as a time when the pressure inthe surge tank 42 is higher than an atmospheric pressure. A time whensupercharging is not performed may be defined as a time when thepressure in the surge tank 42 is lower than or equal to an atmosphericpressure.

In this exemplary configuration, a supercharging system 49 is formed bythe supercharger 44, the bypass passage 47, and the air bypass valve 48.The air bypass valve 48 is one component of the state quantity settingdevice.

The engine 1 has a swirl generation section for generating swirl flow inthe combustion chamber 17. The swirl generation section is implementedby a swirl control valve 56 mounted in the intake passage 40 as shown inFIG. 3. The swirl control valve 56 is disposed in a secondary passage402 among a primary passage 401 connected to the first intake port 181and the secondary passage 402 connected to the second intake port 182.The swirl control valve 56 is an opening degree adjusting valve that canregulate the cross-sectional surface of the secondary passage. When theopening degree of the swirl control valve 56 is small, a flow rate ofintake air flowing into the combustion chamber 17 from the first intakeport 181 among the first intake port 181 and the second intake port 182aligned in the front-rear direction of the engine 1 is relativelyincreased, and a flow rate of intake air flowing into the combustionchamber 17 from the second intake port 182 is relatively reduced, sothat swirl flow in the combustion chamber 17 becomes strong. When theopening degree of the swirl control valve 56 is great, flow rates ofintake air flowing into the combustion chamber 17 from the first intakeport 181 and the second intake port 182, respectively, are almost equal,so that swirl flow in the combustion chamber 17 is weakened. When theswirl control valve 56 is fully opened, no swirl flow is generated. Theswirl flow circulates in the counterclockwise direction in FIG. 3 asindicated by outlined arrows (also see outlined arrows in FIG. 2).

The swirl generation section may be structured such that, instead of orin addition to the swirl control valve 56 being disposed in the intakepassage 40, periods in which two intake valves 21 are opened are shiftedfrom each other, and intake air can be introduced into the combustionchamber 17 from only one of the intake valves 21. By only one of the twointake valves 21 being opened, intake air is non-uniformly introducedinto the combustion chamber 17, so that swirl flow can be generated inthe combustion chamber 17. Furthermore, the swirl generation section maybe structured such that the shape of the intake port 18 is properlydesigned to generate swirl flow in the combustion chamber 17.

The strength of the swirl flow in the combustion chamber 17 is defined.In this exemplary configuration, the strength of the swirl flow in thecombustion chamber 17 is represented by “swirl ratio”. The “swirl ratio”can be defined as a value obtained by dividing, by an engine angularvelocity, a value that is obtained by an intake-air-flow lateral angularvelocity measured for each valve lift being integrated. Theintake-air-flow lateral angular velocity can be obtained based on themeasurement using a rig tester shown in FIG. 5. That is, in the testershown in FIG. 5, the cylinder head 13 is vertically inverted and set ona base, and the intake port 18 is connected to a not-illustrated intakeair supply device. Furthermore, a cylinder 36 is set on the cylinderhead 13, and an impulse meter 38 having a honeycomb-shaped rotor 37 isconnected to the upper end of the cylinder 36. The lower surface of theimpulse meter 38 is positioned so as to be distant, by 1.75 D (Drepresents a cylinder bore diameter), from a mating surface on which thecylinder head 13 and a cylinder block are mated with each other. Atorque that acts on the honeycomb-shaped rotor 37 due to swirl (see anarrow in FIG. 5) generated in the cylinder 36 according to intake airbeing supplied, is measured by the impulse meter 38, and theintake-air-flow lateral angular velocity can be obtained based thereon.

FIG. 6 illustrates a relationship between a swirl ratio and an openingdegree of the swirl control valve 56 in the engine 1. In FIG. 6, theopening degree of the swirl control valve 56 is represented by anopening ratio of opening of the swirl control valve 56 to a fully opencross-section of the secondary passage 402. When the swirl control valve56 is fully closed, the opening ratio with respect to the secondarypassage 402 is 0%. When the opening degree of the swirl control valve 56is increased, the opening ratio with respect to the secondary passage402 is greater than 0%. When the swirl control valve 56 is fully opened,the opening ratio with respect to the secondary passage 402 is 100%. Asillustrated in FIG. 6, in the engine 1, when the swirl control valve 56is fully closed, the swirl ratio is about 6. When the swirl ratio is tobe 4 or greater, the opening degree of the swirl control valve 56 may beadjusted such that the opening ratio is 0 to 15%.

An exhaust passage 50 is connected to the other side surface of theengine 1. The exhaust passage 50 communicates with the exhaust port 19of each cylinder 11. Through the exhaust passage 50, exhaust gasdischarged from the combustion chamber 17 flows. The upstream portion ofthe exhaust passage 50 forms independent passages that diverge for therespective cylinders 11, which is not shown in detail. The upstream endof the independent passage is connected to the exhaust port 19 of eachcylinder 11.

In the exhaust passage 50, an exhaust gas purification system having oneor more catalytic converters is disposed. In this exemplaryconfiguration, the exhaust gas purification system has two catalyticconverters. The upstream-side catalytic converter is disposed in anengine compartment. The upstream-side catalytic converter includes athree-way catalyst 511 and a GPF (Gasoline Particulate Filter) 512. Thedownstream-side catalytic converter is disposed outside the enginecompartment. The downstream-side catalytic converter includes athree-way catalyst 513. The exhaust gas purification system is notlimited to one having the illustrated configuration. For example, theGPF may not be provided. Furthermore, the catalytic converter may notinclude a three-way catalyst. Moreover, the order in which the three-waycatalyst and the GPF are arranged may be changed as appropriate.

An EGR passage 52 that forms an external EGR system is connected betweenthe intake passage 40 and the exhaust passage 50. The EGR passage 52 isa passage for recirculating a part of burned gas into the intake passage40. The upstream end of the EGR passage 52 is connected between theupstream-side catalytic converter and the downstream-side catalyticconverter in the exhaust passage 50. The downstream end of the EGRpassage 52 is connected to a portion upstream of the supercharger 44 inthe intake passage 40. More specifically, the downstream end of the EGRpassage 52 is connected to a mid-portion of the bypass passage 47. EGRgas that flows in the EGR passage 52 flows into the portion upstream ofthe supercharger 44 in the intake passage 40 without passing through theair bypass valve 48 of the bypass passage 47.

A water-cooling type EGR cooler 53 is disposed in the EGR passage 52.The EGR cooler 53 is configured to cool burned gas. An EGR valve 54 isalso disposed in the EGR passage 52. The EGR valve 54 is configured toadjust a flow rate of burned gas that flows in the EGR passage 52. Theopening degree of the EGR valve 54 is adjusted, whereby an amount ofcooled burned gas, that is, external EGR gas, to be recirculated can beadjusted.

In this exemplary configuration, an EGR system 55 is structured by theexternal EGR system that includes the EGR passage 52 and the EGR valve54, and the internal EGR system that includes the intake electric S-VT23 and the exhaust electric S-VT 24 described above. Furthermore, theEGR valve 54 is one component of the state quantity setting device. Inthe external EGR system, the EGR passage 52 is connected to a portiondownstream of the catalytic converter, and the EGR passage 52 has theEGR cooler 53. Therefore, burned gas having a temperature lower thanthat in the internal EGR system can be supplied to the combustionchamber 17.

The control apparatus for the engine includes the ECU (Engine ControlUnit) 10 for operating the engine 1. The ECU 10 is a controller based ona known microcomputer. The ECU 10 includes a central processing unit(CPU) 101 for executing a program, a memory 102 implemented by, forexample, a RAM (Random Access Memory) or a ROM (Read Only Memory) forstoring programs and data, and an input/output bus 103 for inputting andoutputting an electrical signal. The ECU 10 is an example of acontroller.

As shown in FIG. 1 and FIG. 4, various sensors SW1 to SW16 are connectedto the ECU 10. The sensors SW1 to SW16 output detection signals to theECU 10. The sensors include sensors described below.

That is, the sensors include: an air flow sensor SW1, disposeddownstream of the air cleaner 41 in the intake passage 40, for detectinga flow rate of fresh air that flows in the intake passage 40; a firstintake air temperature sensor SW2, disposed downstream of the aircleaner 41 in the intake passage 40, for detecting a temperature offresh air; a first pressure sensor SW3, disposed downstream of aposition at which the EGR passage 52 is connected to the intake passage40 and disposed upstream of the supercharger 44, for detecting pressureof gas that flows into the supercharger 44; a second intake airtemperature sensor SW4, disposed downstream of the supercharger 44 inthe intake passage 40 and disposed upstream of a position at which thebypass passage 47 is connected to the intake passage 40, for detecting atemperature of gas that flows from the supercharger 44; a secondpressure sensor SW5 mounted to the surge tank 42 for detecting pressureof gas flowing downstream of the supercharger 44; pressure sensors SW6,mounted to the cylinder heads 13 so as to correspond to the cylinders11, respectively, each of which detects pressure in the combustionchamber 17; an exhaust air temperature sensor SW7 disposed in theexhaust passage 50 for detecting a temperature of exhaust gas dischargedfrom the combustion chamber 17; a linear 02 sensor SW8, disposedupstream of the upstream-side catalytic converter in the exhaust passage50, for detecting the concentration of oxygen in exhaust gas; a lambda02 sensor SW9, disposed downstream of the three-way catalyst 511 in theupstream-side catalytic converter, for detecting the concentration ofoxygen in exhaust gas; a water temperature sensor SW10 mounted to theengine 1 for detecting a temperature of cooling water; a crank anglesensor SW11 mounted to the engine 1 for detecting a rotational angle ofthe crankshaft 15; an accelerator opening degree sensor SW12 mounted toan accelerator pedal mechanism for detecting an accelerator openingdegree corresponding to an amount of operation of an accelerator pedal;an intake cam angle sensor SW13 mounted to the engine 1 for detecting arotational angle of an intake cam shaft; an exhaust cam angle sensorSW14 mounted to the engine 1 for detecting a rotational angle of anexhaust cam shaft; an EGR differential pressure sensor SW15 disposed inthe EGR passage 52 for detecting a differential pressure between theupstream side and the downstream side of the EGR valve 54; and a fuelpressure sensor SW16, mounted to the common rail 64 of the fuel supplysystem 61, for detecting pressure of fuel to be supplied to the injector6.

The ECU 10 determines an operation state of the engine 1 and calculatesa control amount for each device, based on the detection signals. TheECU 10 outputs control signals based on the calculated control amounts,to the injector 6, the spark plug 25, the intake electric S-VT 23, theexhaust electric S-VT 24, the fuel supply system 61, the throttle valve43, the EGR valve 54, the electromagnetic clutch 45 of the supercharger44, the air bypass valve 48, and the swirl control valve 56.

For example, the ECU 10 sets a target torque for the engine 1 anddetermines a target boost pressure based on a detection signal from theaccelerator opening degree sensor SW12 and a preset map. The ECU 10adjusts an opening degree of the air bypass valve 48 based on the targetboost pressure, and a differential pressure, between the front side andthe rear side of the supercharger 44, obtained according to detectionsignals from the first pressure sensor SW3 and the second pressuresensor SW5, thereby performing feedback control such that the boostpressure becomes the target boost pressure.

Furthermore, the ECU 10 sets a target EGR ratio (that is, ratio of EGRgas to the total gas in the combustion chamber 17) based on theoperation state of the engine 1 and a preset map. The ECU 10 determinesa target EGR gas amount according to the target EGR ratio and an amountof intake air based on a detection signal from the accelerator openingdegree sensor SW12, and adjusts an opening degree of the EGR valve 54based on the differential pressure, between the front side and the rearside of the EGR valve 54, obtained according to a detection signal fromthe EGR differential pressure sensor SW15, thereby performing feedbackcontrol such that an amount of external EGR gas to be introduced intothe combustion chamber 17 becomes the target EGR gas amount.

Moreover, the ECU 10 performs air-fuel ratio feedback control when apredetermined control condition is satisfied. Specifically, the ECU 10adjusts an amount of fuel injected by the injector 6 such that anair-fuel ratio of air-fuel mixture has a desired value, based on theconcentration of oxygen, in exhaust air, detected by the linear O₂sensor SW8 and the lambda O₂ sensor SW9.

The other controls of the engine 1 by the ECU 10 will be described belowin detail.

(Operation Region of Engine)

FIG. 7 illustrates an operation region map 701 for the engine 1 in theupper diagram. The operation region map 701 is defined according toloads and the number of revolutions. The operation region map 701 isdivided into three regions according to whether load is high or low.Specifically, the three regions are a low load region (A) that includesan idling operation region, a high load region (C) that includes fullload, and an intermediate load region (B) between the low load region(A) and the high load region (C). In the operation region map 701,combustion by compression autoignition is performed in the intermediateload region mainly for improving fuel economy and improving exhaust gasperformance in the engine 1. Hereinafter, combustion modes in the lowload region, the intermediate load region, and the high load region,will be sequentially described.

In the following description, FIG. 18 to FIG. 22 are referred to, asappropriate, in the case of describing the ignition timing when thespark plug 25 ignites air-fuel mixture in the combustion chamber 17.Here, FIG. 18 illustrates difference among combustion waveforms when theignition timing is changed in accordance with the initial temperature inthe combustion chamber 17. FIG. 19 illustrates change in the ignitiontiming with respect to the initial temperature in the combustion chamber17, in a region in which the load is a predetermined load or higher in anon-supercharge SPCCI region, and FIG. 20 illustrates change in the SIrate with respect to the ignition timing in a region in which the loadis a predetermined load or higher in the non-supercharge SPCCI region.FIG. 21 illustrates change in the ignition timing with respect to theload on the engine 1, and FIG. 22 illustrates change in the initialtemperature in the combustion chamber 17 with respect to the load on theengine 1, in a region in which the load is a predetermined load orhigher in the non-supercharge SPCCI region.

(Low Load Region)

When the operation state of the engine 1 is in the low load region (thatis, the engine 1 operates with load lower than a first load), thecombustion mode is SI combustion in which the spark plug 25 ignitesair-fuel mixture in the combustion chamber 17 and the air-fuel mixtureis thus combusted by flame propagation. This is because assuredlyobtaining of combustion stability is prioritized. Hereinafter, thecombustion mode in the low load region may be referred to as low load SIcombustion.

When the operation state of the engine 1 is in the low load region, anair-fuel ratio (A/F) of air-fuel mixture is the theoretical air-fuelratio (A/F≈14.7). In the following description, values of an air-fuelratio of air-fuel mixture, an excess air ratio K, and a G/F representvalues at ignition timing. When the air-fuel ratio of air-fuel mixtureis the theoretical air-fuel ratio, exhaust gas discharged from thecombustion chamber 17 can be purified by the three-way catalyst, so thatthe engine 1 has good exhaust gas performance. The A/F of air-fuelmixture may be set so as to fall within a purification window of thethree-way catalyst. The excess air ratio K of the air-fuel mixture maybe 1.0±0.2.

When the operation state of the engine 1 is in the low load region, theEGR system 55 introduces EGR gas into the combustion chamber 17 in orderto improve fuel economy performance of the engine 1. The G/F of theair-fuel mixture, that is, a weight ratio between the total gas and fuelin the combustion chamber 17 is set in a range from 18 to 30. The G/F ofthe air-fuel mixture may be set in a range from 18 to 50. The air-fuelmixture is EGR gas lean. The dilution rate of the air-fuel mixture ishigh. When the G/F of the air-fuel mixture is, for example, 25, SIcombustion can be stably performed without causing autoignition of theair-fuel mixture, in the low load operation region. In the low loadregion, the G/F of the air-fuel mixture is almost uniformly maintainedregardless of whether or not load on the engine 1 is high or low. Thus,the SI combustion is stabilized over the entirety of the low loadregion. Furthermore, fuel economy of the engine 1 is improved, andexhaust gas performance becomes good.

When the operation state of the engine 1 is in the low load region, anamount of fuel is small. Therefore, in order to set λ of the air-fuelmixture to 1.0±0.2 and set the G/F in a range from 18 to 50, an amountof gas to be filled in the combustion chamber 17 needs to be less than100%. Specifically, the engine 1 performs throttling for adjusting anopening degree of the throttle valve 43 and/or Miller cycle fordelaying, to the intake bottom dead center or later, the closing time ofthe intake valve 21.

Furthermore, when the operation state of the engine 1 is in the low loadregion, an opening degree of the swirl control valve 56 is adjusted asappropriate.

In the low-load low-rotation region in the low load region, an amount offilled gas is further reduced, whereby the combustion temperature of theair-fuel mixture and the temperature of exhaust gas may be enhanced.Thus, the catalytic converter is advantageously maintained in an activestate.

(Intermediate Load Region)

When the operation state of the engine 1 is in the intermediate loadregion, an amount of injected fuel is increased. The temperature in thecombustion chamber 17 becomes high, whereby autoignition can be stablyperformed. The engine 1 performs CI combustion in the intermediate loadregion in order to improve fuel economy and improve exhaust gasperformance.

In combustion by autoignition, if the temperature in the combustionchamber 17 before start of compression varies, timing of autoignition isgreatly changed. Therefore, in the intermediate load region, the engine1 performs SPCCI combustion in which SI combustion and CI combustion arecombined. In the SPCCI combustion, the spark plug 25 forcibly ignitesair-fuel mixture in the combustion chamber 17, and the air-fuel mixtureis thus combusted by flame propagation, and unburned air-fuel mixture iscombusted by autoignition due to heat generation by the SI combustionenhancing the temperature in the combustion chamber 17. By adjusting anamount of heat generated by the SI combustion, variation of thetemperature in the combustion chamber 17 before start of compression canbe absorbed.

In details, even if the temperature in the combustion chamber 17 beforestart of compression varies, when the start timing of the SI combustionis adjusted by, for example, adjustment of ignition timing, autoignitionof unburned air-fuel mixture can be performed at target timing. That is,as described above, in combustion by autoignition, the autoignitiontiming changes in accordance with the temperature in the combustionchamber 17 before start of compression. If the autoignition timingbecomes earlier, start of CI combustion is made earlier, whereby SIcombustion is shortened accordingly. As a result, combustion by flamepropagation is not sufficiently performed in the combustion chamber, andCI combustion becomes steep. This might pose an obstacle to suppressionof occurrence of combustion noise. On the other hand, if theautoignition timing is delayed, stability of CI combustion might belost, leading to increase in unburned fuel or reduction in exhaust gasperformance. Hereinafter, the temperature in the combustion chamber 17before start of compression may be referred to as “initial temperature”.

Accordingly, if start timing of SI combustion is adjusted through changeof the ignition timing, the autoignition timing can be controlled evenif the initial temperature in the combustion chamber 17 varies.

Specifically, in the example shown in FIG. 18, under the assumption thatthe load on the engine 1 is the same, comparison is made among anignition timing tIg1 when the initial temperature in the combustionchamber 17 is a first temperature, an ignition timing tIg2 when theinitial temperature in the combustion chamber 17 is a secondtemperature, an ignition timing tIg3 when the initial temperature in thecombustion chamber 17 is a third temperature, and an ignition timingtIg4 when the initial temperature in the combustion chamber 17 is afourth temperature. It is noted that the first to fourth temperatureshave a relationship of fourth temperature <third temperature <secondtemperature <first temperature.

In non-supercharge SPCCI combustion, even under the same load, theignition timing is advanced as the initial temperature in the combustionchamber 17 decreases, as shown in FIG. 18 and FIG. 19. If the ignitiontiming is advanced, start timing of SI combustion becomes earlier, sothat the amount of heat generation by SI combustion increases and, asshown in FIG. 20, the SI rate increases. In this way, if the SI rate isadjusted in accordance with the initial temperature in the combustionchamber 17, variation in the initial temperature in the combustionchamber 17 is compensated for by heat generation by flame propagation,and thus, in non-supercharge SPCCI combustion, autoignition timings canbe equalized at a desired timing tCI.

Thus, in the SPCCI combustion, in order to accurately control the timingof autoignition, the timing of autoignition needs to change according tothe ignition timing being changed. Sensitivity for change of timing ofautoignition with respect to change of ignition timing is preferablyhigh.

The inventors of the present invention have found through study that,when a G/F of the air-fuel mixture is in a range from 18 to 50, theSPCCI combustion can be stably performed and timing of autoignitionchanges with respect to change of ignition timing. When the operationstate of the engine 1 is in the intermediate load region, the engine 1sets a state inside the combustion chamber 17 such that λ of air-fuelmixture is 1.0±0.2 and a G/F of the air-fuel mixture is in a range from18 to 50.

Furthermore, the engine 1 adjusts the opening degree of the swirlcontrol valve 56 as appropriate. When the operation state of the engine1 is in the intermediate load region, the swirl control valve 56 isfully closed or opened to a predetermined opening degree on the closingside. In the combustion chamber 17, relatively strong swirl flow isformed. At the timing of ignition, the swirl ratio may be 4 or greater.

By timing of autoignition being accurately controlled in the SPCCIcombustion, increase of combustion noise can be avoided when theoperation state of the engine 1 is in the intermediate load region.Furthermore, a dilution rate of the air-fuel mixture is set to be ashigh as possible to perform CI combustion, whereby the engine 1 can havehigh fuel economy performance. Moreover, λ of air-fuel mixture is set to1.0±0.2, whereby exhaust gas can be purified by the three-way catalyst,so that the engine 1 has good exhaust gas performance.

As described above, in the low load region, a G/F of air-fuel mixture isin a range from 18 to 50 (for example, 25), and λ of the air-fuelmixture is set to 1.0±0.2. The state quantity of the combustion chamber17 does not greatly change between when the operation state of theengine 1 is in the low load region and when the operation state of theengine 1 is in the intermediate load region. Therefore, robustness ofcontrol of the engine 1 with respect to change of load on the engine 1is enhanced.

When the operation state of the engine 1 is in the intermediate loadregion, an amount of fuel is increased unlike in the low load region.Therefore, an amount of gas to be filled in the combustion chamber 17need not be adjusted. The opening degree of the throttle valve 43 is afully open degree.

When load on the engine 1 is increased and an amount of fuel is furtherincreased, an amount of gas introduced into the combustion chamber 17 bynatural aspiration is insufficient for setting λ of air-fuel mixture to1.0±0.2 and setting a G/F of the air-fuel mixture in a range from 18 to50. Therefore, the supercharger 44 supercharges the combustion chamber17 with gas to be introduced thereinto, in a region in which load ishigher than a predetermined load (that is, third load) in theintermediate load region. The intermediate load region (B) is dividedinto a first intermediate load region (B1) in which the load is higherthan a predetermined load and supercharging is performed, and a secondintermediate load region (B2) in which the load is the predeterminedload or lower load and no supercharging is performed. The predeterminedload is, for example, ½ load. The second intermediate load region is aregion in which load is lower than load in the first intermediate loadregion. Hereinafter, the combustion mode in the first intermediate loadregion may be referred to as supercharge SPCCI combustion, and thecombustion mode in the second intermediate load region may be referredto as non-supercharge SPCCI combustion.

In the second intermediate load region in which no supercharging isperformed, as an amount of fuel is increased, fresh air to be introducedinto the combustion chamber 17 is increased, while EGR gas is reduced.The G/F of the air-fuel mixture is reduced when load on the engine 1 isincreased. Since the opening degree of the throttle valve 43 is a fullyopen degree, the engine 1 adjusts an amount of EGR gas to be introducedinto the combustion chamber 17, whereby an amount of fresh air to beintroduced into the combustion chamber 17 is adjusted. In the secondintermediate load region, the state quantity of the combustion chamber17 is such that, for example, λ of the air-fuel mixture is almostuniformly 1.0, while the G/F of the air-fuel mixture changes in a rangefrom 25 to 28.

Meanwhile, in the first intermediate load region in which superchargingis performed, as an amount of fuel is increased, both fresh air and EGRgas to be introduced into the combustion chamber 17 are increased, inthe engine 1. The G/F of the air-fuel mixture is almost constant evenwhen the load on the engine 1 is increased. In the first intermediateload region, the state quantity of the combustion chamber 17 is suchthat, for example, λ of the air-fuel mixture is almost uniformly 1.0 andthe G/F of the air-fuel mixture is almost uniformly 25.

As described later, in the case of performing SPCCI combustion, the ECU10 adjusts operation quantity relevant to the initial temperature in thecombustion chamber 17 such that, when the load is high (in thisconfiguration example, when the operation state of the engine 1 is inthe first intermediate load region (B1)), the initial temperaturebecomes lower than when the load is low (in this configuration example,when the operation state of the engine 1 is in the second intermediateload region (B2)). In this configuration example, the operation quantitycorresponds to internal EGR gas and external EGR gas.

In the case of performing SPCCI combustion, when the load is high (firstintermediate load region (B1)), the ECU 10 advances the ignition timingthan when the load is low (second intermediate load region (B2)),thereby increasing the SI rate.

In the case where the operation state of the engine 1 is in the secondintermediate load region (B2), the ECU 10 adjusts the operation quantitysuch that the temperature in the combustion chamber 16 before start ofcompression decreases as the load on the engine 1 increases, therebyincreasing the SI rate.

In the case where the operation state of the engine 1 is in the secondintermediate load region (B2), as the load on the engine 1 increases,the ECU 10 advances the ignition timing, thereby increasing the SI rate.

In the case where the operation state of the engine 1 is in the firstintermediate load region (B1), as the load on the engine 1 increases,the ECU 10 advances the ignition timing, thereby making the SI rateconstant with respect to change in the load on the engine 1.

(High Load Region)

The combustion mode is SI combustion when the operation state of theengine 1 is in the high load region. This is because assuredly avoidingof combustion noise is prioritized. Hereinafter, the combustion mode inthe high load region may be referred to as high load SI combustion.

When the operation state of the engine 1 is in the high load region, λof the air-fuel mixture is 1.0±0.2. Furthermore, the G/F of the air-fuelmixture is set in a range from 18 to 30. The G/F of the air-fuel mixturemay be set in a range from 18 to 50. In the high load region, theopening degree of the throttle valve 43 is a fully open degree, and thesupercharger 44 performs supercharging.

In the high load region, an amount of EGR gas is reduced according toload being increased in the engine 1. The G/F of the air-fuel mixturebecomes small when the load on the engine 1 is increased. An amount offresh air to be introduced into the combustion chamber 17 is increasedaccording to an amount of the EGR gas being reduced. Therefore, anamount of fuel can be increased. The maximum output of the engine 1 isadvantageously enhanced.

Furthermore, the engine 1 adjusts an opening degree of the swirl controlvalve 56 as appropriate.

The state quantity of the combustion chamber 17 does not greatly changebetween when the operation state of the engine 1 is in the high loadregion and when the operation state of the engine 1 is the intermediateload region. Robustness of control of the engine 1 with respect tochange of load on the engine 1 is enhanced.

As described above, the engine 1 performs SI combustion in the high loadregion. However, a problem arises that abnormal combustion such aspreignition or knocking is likely to occur.

Therefore, the engine 1 is configured to avoid the abnormal combustionby fuel injection mode being properly designed for the high load region.Specifically, the ECU 10 outputs control signals to the fuel supplysystem 61 and the injector 6 such that fuel is injected into thecombustion chamber 17 with high fuel pressure of 30 MPa or higher attiming in a period (hereinafter, the period is referred to as retardperiod) between the later stage of the compression stroke and theinitial stage of the expansion stroke. The ECU 10 further outputs acontrol signal to the spark plug 25 such that the air-fuel mixture isignited at timing at or near the compression top dead center afterinjection of fuel. In the below description, injection of fuel into thecombustion chamber 17 with high fuel pressure at timing in the retardperiod is referred to as high-pressure retard injection.

In the high-pressure retard injection, a time for reaction of theair-fuel mixture is shortened to avoid abnormal combustion. That is, thetime for reaction of the air-fuel mixture is a time obtained by additionof (1) a period (that is, injection period) in which the injector 6injects fuel, (2) a period (that is, air-fuel mixture forming period) inwhich burnable air-fuel mixture is formed around the spark plug 25 afterinjection of fuel has ended, and (3) a period (that is, combustionperiod) up to the end of the SI combustion started by ignition.

When fuel is injected into the combustion chamber 17 with a high fuelpressure, the injection period and the air-fuel mixture forming periodare each shortened. When the injection period and the air-fuel mixtureforming period are shortened, timing at which injection of fuel startscan be close to ignition timing. In the high-pressure retard injection,fuel is injected into the combustion chamber 17 with a high pressure.Therefore, fuel injection is performed at timing in the retard periodfrom the later stage of the compression stroke to the initial stage ofthe expansion stroke.

When fuel is injected into the combustion chamber 17 with a high fuelpressure, turbulent energy in the combustion chamber 17 becomes high.When the timing of fuel injection is made close to the compression topdead center, the SI combustion can be started in a state where theturbulent energy in the combustion chamber 17 is high. As a result,combustion period is shortened.

The high-pressure retard injection allows the injection period, theair-fuel mixture forming period, and the combustion period to be eachshortened. As compared to a case where fuel is injected into thecombustion chamber 17 in the intake stroke, the high-pressure retardinjection allows a time for reaction of the air-fuel mixture to begreatly shortened. The high-pressure retard injection allows a time forreaction of the air-fuel mixture to be shortened, whereby abnormalcombustion can be avoided.

In the technical field of the engine control, retarding of ignitiontiming has been conventionally performed in order to avoid abnormalcombustion. However, when the ignition timing is retarded, fuel economyperformance is degraded. If the high-pressure retard injection isperformed, it is possible to suppress the degree of retarding of theignition timing. By using the high-pressure retard injection, the fueleconomy performance is improved.

When the fuel pressure is, for example, 30 MPa or higher, the injectionperiod, the air-fuel mixture forming period, and the combustion periodcan be effectively shortened. The fuel pressure may be preferably set asappropriate according to the properties of fuel. The upper limit valueof the fuel pressure may be, for example, 120 MPa.

When the number of revolutions of the engine 1 is small, time for whicha crank angle is changed by the same angle is long. Therefore,shortening of a time in which air-fuel mixture can react, by thehigh-pressure retard injection, is particularly effective for avoidingabnormal combustion. Meanwhile, when the number of revolutions of theengine 1 is great, time for which a crank angle is changed by the sameangle is short. Therefore, shortening of a time in which air-fuelmixture can react is not so effective for avoiding abnormal combustion.

Furthermore, in the high-pressure retard injection, injection of fuelinto the combustion chamber 17 starts around the compression top deadcenter. Therefore, in the compression stroke, gas which does not containfuel, in other words, gas having a high specific heat ratio iscompressed in the combustion chamber 17. In a case where the number ofrevolutions of the engine 1 is great, when the high-pressure retardinjection is performed, the temperature in the combustion chamber 17 atthe compression top dead center, that is, the compression endtemperature becomes high. When the compression end temperature becomeshigh, abnormal combustion such as knocking may be caused.

In the engine 1, the high load region (C) is divided into a first highload region (C1) on the low rotation side, and a second high load region(C2) in which the number of revolutions is higher than that in the firsthigh load region (C1). When the high load region is divided into threeequal regions that are a low rotation region, an intermediate rotationregion, and a high rotation region, the first high load region mayinclude the low rotation region and the intermediate rotation region.When the high load region is divided into the three equal regions thatare the low rotation region, the intermediate rotation region, and thehigh rotation region, the second high load region may include the highrotation region.

In the first high load region, the injector 6 receives a control signalfrom the ECU 10, to perform the above-described high-pressure retardinjection. In the second high load region, the injector 6 receives acontrol signal from the ECU 10, to perform fuel injection atpredetermined timing in the intake stroke. The fuel injection in theintake stroke does not require a high fuel pressure. The ECU 10 outputsa control signal to the fuel supply system 61 such that the fuelpressure is lower than a fuel pressure in the high-pressure retardinjection (for example, such that the fuel pressure is less than 40MPa). By the fuel pressure being lowered, the mechanical resistance lossof the engine 1 is reduced, so that fuel economy is advantageouslyimproved.

When fuel is injected into the combustion chamber 17 in the intakestroke, the specific heat ratio of gas in the combustion chamber 17 isreduced, whereby the compression end temperature becomes low. Thecompression end temperature becomes low, whereby the engine 1 is allowedto avoid abnormal combustion. The ignition timing need not be retardedin order to avoid abnormal combustion, whereby the spark plug 25 ignitesair-fuel mixture at timing of or at timing near the compression top deadcenter in the second high load region, as in the first high load region.

In the first high load region, the high-pressure retard injectionprevents occurrence of autoignition of the air-fuel mixture. Therefore,the engine 1 can perform stable SI combustion. In the second high loadregion, the fuel injection in the intake stroke prevents occurrence ofautoignition of the air-fuel mixture. Therefore, the engine 1 canperform stable SI combustion.

(SPCCI Combustion)

The SPCCI combustion will be described. FIG. 8 illustrates, in the upperdiagram, a waveform 801 representing an example of change of a heatgeneration rate with respect to a crank angle in the SPCCI combustion.When the spark plug 25 ignites air-fuel mixture around the compressiontop dead center, more accurately, at predetermined timing before thecompression top dead center, combustion by flame propagation starts.Heat generation at SI combustion is gentler than heat generation at CIcombustion. Therefore, the waveform of the heat generation rate has arelatively small slope. A pressure variation (dp/dθ) in the combustionchamber 17 at the SI combustion is gentler than that at the CIcombustion, which is not shown.

When the temperature and pressure in the combustion chamber 17 areenhanced by the SI combustion, autoignition of unburned air-fuel mixtureoccurs. In the example of the waveform 801, the slope of the waveform ofthe heat generation rate changes from small to large at or near thecompression top dead center. That is, the waveform of the heatgeneration rate has an inflection point at timing when the CI combustionstarts.

After the start of the CI combustion, the SI combustion and the CIcombustion are performed in parallel. In CI combustion, heat generationis greater than in the SI combustion. Therefore, the heat generationrate becomes relatively great. However, the CI combustion is performedafter the compression top dead center, and the piston 3 has been moveddownward due to motoring. The slope of the waveform of the heatgeneration rate is prevented from becoming excessively great due to theCI combustion. dp/dθ at the CI combustion becomes relatively gentle.

dp/dθ can be used as an index representing combustion noise. Asdescribed above, dp/dθ can be reduced in the SPCCI combustion, so thatthe combustion noise can be prevented from becoming excessively great.Combustion noise can be reduced to an allowable level or lower level.

When the CI combustion ends, the SPCCI combustion ends. The combustionperiod of the CI combustion is shorter than that of the SI combustion.The combustion end time of the SPCCI combustion is advanced as comparedto the SI combustion. In other words, in the SPCCI combustion, thecombustion end time in the expansion stroke can be made close to thecompression top dead center. The SPCCI combustion is more advantageousthan the SI combustion in improvement of fuel economy performance of theengine 1.

Therefore, the SPCCI combustion enables both prevention of combustionnoise and improvement of fuel economy performance.

An SI rate is defined as a parameter representing the characteristics ofthe SPCCI combustion. The SI rate is defined as an index that isassociated with a ratio of an amount of heat generated by the SIcombustion, to the total amount of heat generated by the SPCCIcombustion. The SI rate is a ratio between amounts of heat generated bytwo combustions in different combustion modes. The SI rate may be aratio of an amount of heat generated by the SI combustion to an amountof heat generated by the SPCCI combustion. For example, in the waveform801, the SI rate can be represented as SI rate=(area of SIcombustion)/(area of SPCCI combustion). In the waveform 801, the SI ratemay be referred to as SI fuel rate to represent a rate of fuel which iscombusted in the SI combustion.

The SI rate is a ratio between SI combustion and CI combustion in theSPCCI combustion in which the SI combustion and the CI combustion arecombined. When the SI rate is high, the proportion of the SI combustionis high. When the SI rate is low, the proportion of the CI combustion ishigh.

The SI rate is not limited to one defined as described above. The SIrate may be variously defined. For example, the SI rate may be a ratioof an amount of heat generated by the SI combustion, to an amount ofheat generated by the CI combustion. That is, in the waveform 801, theSI rate=(area of SI combustion)/(area of CI combustion) may besatisfied.

Furthermore, in the SPCCI combustion, a waveform of the heat generationrate has an inflection point at timing when the CI combustion starts. Inthe intermediate diagram of FIG. 8, as indicated by reference numeral802, the inflection point in the waveform of the heat generation rate isset as a boundary, and a range on the advance side relative to theboundary may be defined as the SI combustion, and a range on the retardside relative to the boundary may be defined as the CI combustion. Inthis case, the SI rate may be represented as SIrate=Q_(SI)/(Q_(SI)+Q_(CI)) or SI rate=Q_(SI)/Q_(CI) based on an areaQ_(SI) of the range on the advance side relative to the boundary and anarea Q_(CI) of the range on the retard side relative to the boundary, asindicated by hatching in the wavelength 802. Furthermore, the SI ratemay be defined based on not the entirety but a part of the area of therange on the advance side relative to the boundary, and a part of thearea of the range on the retard side relative to the boundary.

Furthermore, the SI rate may not be defined based on heat generation.The SI rate may be represented as SI rate=Δθ_(SI)/(Δθ_(SI)+Δθ_(CI)) orSI rate=Δθ_(SI)/Δθ_(CI) based on a crank angle Δθ_(SI) for the range onthe advance side relative to the boundary and a crank angle Δθ_(CI) forthe range on the retard side relative to the boundary.

Moreover, SI rate=ΔP_(SI)/(ΔP_(SI)+ΔP_(CI)) or SI rate=ΔP_(SI)/ΔP_(CI)may be defined based on a peak ΔP_(SI) of the heat generation rate inthe range on the advance side relative to the boundary and a peakΔP_(CI) of the heat generation rate in the range on the retard siderelative to the boundary.

In addition, SI rate=φ_(SI)/(φ_(SI)+φ_(CI)) or SI rate=φ_(SI)/φ_(CI) maybe defined based on a slope φ_(SI) of the heat generation rate in therange on the advance side relative to the boundary and a slope φ_(CI) ofthe heat generation rate in the range on the retard side relative to theboundary.

In the description herein, the SI rate is defined, based on the waveformof the heat generation rate, from an area (that is, magnitude of anamount of generated heat), the length of the horizontal axis (that is,magnitude of crank angle), the length of the vertical axis (that is,magnitude of heat generation rate), or a slope (that is, change rate ofheat generation rate). The SI rate may be defined based on a waveform ofa pressure (P) in the combustion chamber 17, similarly from the area,the length of the horizontal axis, the length of the vertical axis, orthe slope, which is not shown.

Furthermore, in the SPCCI combustion, the inflection point of thecombustion waveform associated with the heat generation rate or pressuremay not always appear clearly. The SI rate which is not based on theinflection point may be defined as follows. That is, in the lowerdiagram in FIG. 8, as indicated by reference numeral 803, in thecombustion waveform, a range on the advance side relative to thecompression top dead center (TDC) may be defined as the SI combustion,and a range on the retard side relative to the compression top deadcenter may be defined as the CI combustion. Based thereon, similarly asdescribed above, the SI rate may be defined from the area (Q_(SI),Q_(CI)), the length of the horizontal axis (Δθ_(SI), Δθ_(CI)), thelength of the vertical axis (ΔP_(SI), ΔP_(CI)), or the slope (φ_(SI),φ_(CI)).

Furthermore, the SI rate may be defined based on not a combustionwaveform actually obtained in the combustion chamber 17 but an amount offuel. As described below, in the intermediate load region in which theSPCCI combustion is performed, divided injection that includes precedinginjection and succeeding injection may be performed. Fuel injected intothe combustion chamber 17 by the succeeding injection does not diffusein the combustion chamber 17 and is positioned near the spark plug 25since a time from the injection to ignition is short. Therefore, thefuel injected into the combustion chamber 17 by the succeeding injectionis combusted mainly by the SI combustion. Meanwhile, fuel injected intothe combustion chamber 17 by the preceding injection is combusted mainlyby CI combustion. Therefore, the SI rate can be defined based on anamount of fuel (m₁) injected by the preceding injection, and an amountof fuel (m₂) injected by the succeeding injection. That is, SIrate=m₂/(m₁+m₂) may be defined or SI rate=m₂/m₁ may be defined.

(Stabilization of SPCCI Combustion)

A condition for stably performing the SPCCI combustion will bedescribed. The inventors of the present invention have newly foundthrough study that the SI combustion by the flame propagation needs tobe stabilized before autoignition of air-fuel mixture occurs in order toappropriately perform the SPCCI combustion. When the SI combustion isunstable, the entirety of the combustion including the CI combustion isnot stabilized.

One of factors associated with stability of the SI combustion is aturbulent combustion speed. When the turbulent combustion speed is high,the SI combustion is stabilized. The turbulent combustion speed isinfluenced by an air-fuel ratio (or excess air ratio λ) of air-fuelmixture, an EGR ratio (that is, dilution rate) of the air-fuel mixture,a temperature and a pressure in the combustion chamber, turbulent energyin the combustion chamber, or the like.

The inventors of the present invention have examined, throughsimulation, an excess air ratio λ of air-fuel mixture, a dilution rate(in the description herein, G/F that is a weight ratio between the totalgas and fuel in the combustion chamber) of the air-fuel mixture, atemperature and a pressure in the combustion chamber, and turbulentenergy in the combustion chamber, for obtaining a turbulent combustionspeed necessary for assuring stability of the SI combustion. Thecondition for the simulation is a condition that the engine operateswith a low load, and only internal EGR gas is introduced into thecombustion chamber, to make the temperature in the combustion chamber ashigh as possible.

From the viewpoint of assuredly avoiding great combustion noise causedby knocking, the lower limit of the G/F of air-fuel mixture is 18. Thatis, when the G/F of air-fuel mixture is reduced, even if autoignition ofunburned air-fuel mixture occurs, this phenomenon can be regarded asknocking. Furthermore, when a three-way catalyst is used for combustingsuch a lean air-fuel mixture in order to prevent NOx emissions, theexcess air ratio K of the air-fuel mixture is 1.0±0.2.

From the viewpoint of enhancing fuel economy performance of the engine,the G/F of air-fuel mixture is preferably great. Therefore, theinventors of the present invention have examined a relationship betweenthe G/F of air-fuel mixture and turbulent energy necessary for obtaininga desired turbulent combustion speed as shown in the upper diagram ofFIG. 13 (graph of reference numeral 1301). The engine operates at 2000rpm as the number of revolutions with a low load. Furthermore, internalEGR gas is introduced into the combustion chamber. The closing time ofthe intake valve is 91° ABDC. The geometrical compression ratio of theengine is 18.

According to graph 1301, when of the air-fuel mixture is 1.2, thecharacteristic line of the G/F is like a saturation curve representingsaturation at about 30. Meanwhile, when the number of revolutions of theengine is 2000 rpm, turbulent energy of 40 m²/s² can be obtained. It hasbeen newly found that, even if the turbulent energy is obtained so as tobe greater than 40 m²/s², the G/F of the air-fuel mixture is not likelyto be greater than 30. According to graph 1301, the upper limit of theG/F of air-fuel mixture is 30 in order to assure stability of the SIcombustion.

According to the above-described examination, the G/F of air-fuelmixture needs to be set in a range from 18 to 30. According to graph1301, when λ of the air-fuel mixture is 1.0 or 1.2, and the range of theG/F is in a range from 18 to 30, the range of the turbulent energynecessary for stabilizing the SI combustion is 17 to 40 m²/s².

FIG. 13 illustrates, in the intermediate diagram, a relationship,between a temperature at ignition timing and the G/F of air-fuel mixturein the combustion chamber, which is necessary for obtaining a desiredturbulent combustion speed (graph of reference numeral 1302), under thesame condition as in graph 1301. According to graph 1302, when λ of theair-fuel mixture is 1.0 or 1.2, and the G/F is in a range from 18 to 30,a necessary temperature TIg (K) in the combustion chamber at ignitiontiming is 570 to 800 K.

FIG. 13 illustrates, in the lower diagram, a relationship, betweenpressure at ignition timing and the G/F of air-fuel mixture in thecombustion chamber, which is necessary for obtaining a desired turbulentcombustion speed (graph of reference numeral 1303), under the samecondition as in graph 1301. According to graph 1303, when K of theair-fuel mixture is 1.0 or 1.2, and the G/F is in a range from 18 to 30,a necessary pressure PIg (kPa) in the combustion chamber at ignitiontiming is 400 to 920 kPa.

Even if the geometrical compression ratio of the engine is changed in arange of 13 to 20, influence on a relationship between the G/F ofair-fuel mixture, and turbulent energy which is necessary for obtaininga desired turbulent combustion speed, is little, which is not shown.

FIG. 13 illustrates a result of simulation in the case of the number ofrevolutions of the engine being 2000 rpm. When the number of revolutionsof the engine is great, flow of gas in the combustion chamber becomesstrong, so that a desired turbulent combustion speed is likely to beobtained. The ranges of numerical values of the G/F of air-fuel mixture,and the necessary temperature TIg and the necessary pressure PIg incombustion chamber, as described above, are not limited for a specificoperation state of the engine.

In the SPCCI combustion, as described above, timing of autoignition iscontrolled by SI combustion. Timing of autoignition needs to changeaccording to ignition timing being changed in order to accuratelycontrol timing of autoignition such that autoignition of unburnedair-fuel mixture occurs at target timing. Sensitivity for change oftiming of autoignition with respect to change of ignition timing, ispreferably high.

FIG. 14 is a contour view 1401 showing a change rate of change ofautoignition timing with respect to change of ignition timing (=(changeof crank angle at autoignition timing)/(change of crank angle atignition timing)), which is obtained by an experiment. The change raterepresents a magnitude of change, of a crank angle at autoignitiontiming, obtained when ignition timing is changed by 1° as a crank angle.The greater the value of the change rate is, the higher the sensitivityfor change of timing of autoignition with respect to change of ignitiontiming is. The less the value of the change rate is, the lower thesensitivity for change of timing of autoignition with respect to changeof ignition timing is.

In contour view 1401, the vertical axis represents an EGR ratio ofair-fuel mixture, and the horizontal axis represents an A/F of air-fuelmixture. Sensitivity for change of timing of autoignition with respectto change of ignition timing is reduced toward the upper right portionin the view. Sensitivity for change of timing of autoignition isenhanced toward the lower left portion in the view. In the contour view1401, it is found that, in a range, enclosed by a broken line, in whichλ of the air-fuel mixture is 1.0±0.2 and the G/F is in a range from 18to 30, timing of autoignition is changed, with high sensitivity, withrespect to change of ignition timing. The upper limit of the EGR ratiois preferably 40% from the viewpoint of combustion stability.

That is, when a state inside the combustion chamber is such that K ofair-fuel mixture is 1.0±0.2 and the G/F is in a range from 18 to 30, SIcombustion is stabilized, so that autoignition of unburned air-fuelmixture can be caused to accurately occur at target timing in the SPCCIcombustion.

In the above-described examination, the G/F of air-fuel mixture is 30 atthe maximum. Meanwhile, the inventors of the present invention haveconsidered that a dilution rate of air-fuel mixture is enhanced in orderto further improve fuel economy performance.

The inventors of the present invention have paid attention tostratifying of the G/F of air-fuel mixture in the combustion chambersince SPCCI combustion is a combination of SI combustion and CIcombustion. That is, SI combustion in SPCCI combustion is combustion ofair-fuel mixture ignited by the spark plug 25. The air-fuel mixture nearthe spark plug 25 is combusted mainly by SI combustion. Meanwhile, CIcombustion in SPCCI combustion is combustion of unburned air-fuelmixture by autoignition after start of SI combustion. The surroundingair-fuel mixture, which is distant from the spark plug 25, is combustedmainly by CI combustion.

For example, when strong swirl flow is generated in the combustionchamber 17, residual gas (that is, burned gas) in the cavity 31 in thetop surface of the piston 3 can be expelled to the outside of the cavity31. When combustion is almost uniformly distributed over the entirety ofthe combustion chamber 17, the G/F of air-fuel mixture near the sparkplug 25 becomes relatively small according to residual gas being not inthe cavity 31, and the G/F of surrounding air-fuel mixture which isdistant from the spark plug 25 becomes relatively great according toresidual gas being contained. The G/F of air-fuel mixture in thecombustion chamber 17 can be stratified.

The inventors of the present invention have examined, throughsimulation, a condition in which the SPCCI combustion is stabilized in astate where the G/F of air-fuel mixture is stratified, according to theprocedure shown in FIG. 15. In the simulation, as schematicallyindicated by reference numeral 1506 in FIG. 15, the combustion chamber17 is imaginarily sectioned into an SI part which is near the spark plug25 and a CI part around the SI part, and the G/F of air-fuel mixture isdefined for each of the SI part and the CI part. Each of the excess airratio λ of air-fuel mixture of the SI part and the excess air ratio λ ofair-fuel mixture of the CI part is 1. Furthermore, the G/F of the SIpart is less than the G/F of the CI part.

In the simulation, the inventors of the present invention have firstlyassumed the following three constraint conditions (1) to (3) asconstraint conditions for stably performing the SPCCI combustion.

Constraint condition (1): according to a result from the upper diagramin FIG. 13, a condition that the G/F of air-fuel mixture of the SI partis 22 or less is set as the condition for stably performing SIcombustion. The SI part corresponds to air-fuel mixture that does notcontain residual gas and contains external EGR gas as described above.Therefore, the constraint condition (1) that the G/F of air-fuel mixtureof the SI part is 22 or less can be, in other words, a condition thatthe external EGR ratio is 34% or less.

Constraint condition (2): the G/F of air-fuel mixture of the CI part isless than 100, and a temperature of the CI part is higher than 1050 Kwhen the compression top dead center is reached in the combustionchamber. As compared to SI combustion, combustion is stabilized in CIcombustion even if the air-fuel mixture is diluted. However, there is alimitation on the dilution rate of air-fuel mixture. According to theexperiment performed by the inventors of the present invention, when theG/F of air-fuel mixture is less than 100, a desired combustion stabilitycan be assured in a state where the center of gravity of combustion isin a range from the compression top dead center to 10° CA after the topdead center, in the CI combustion.

Furthermore, the temperature, at autoignition, of fuel that containsgasoline is 1050 K in general. This has been confirmed in the experimentperformed by the inventors of the present invention. Therefore, when theG/F of air-fuel mixture of the CI part is less than 100, and thetemperature of the CI part at the compression top dead center is higherthan 1050 K, CI combustion can be stabilized.

Constraint condition (3): since the SPCCI combustion is a combination ofSI combustion and CI combustion, when the proportion of the SIcombustion is great, the proportion of the CI combustion becomes small,and, when the proportion of the SI combustion is small, the proportionof the CI combustion becomes great. When the proportion of the SIcombustion becomes excessively small, the proportion of the CIcombustion becomes excessively great, resulting in combustion noisebeing increased. The proportion (that is, the above-described SI rate)of SI combustion to SPCCI combustion is represented by “SI fuel rate” asa rate of fuel combusted by SI combustion. When only SI combustion isperformed in the SPCCI combustion, the SI fuel rate is 1. In the SPCCIcombustion, according to the proportion of SI combustion being reduced,the SI fuel rate is gradually reduced to be less than 1.

FIG. 15 illustrates, in a graph of reference numeral 1504, a region inwhich combustion noise indicates an allowable value or less in the SPCCIcombustion, and a region in which combustion noise is greater than theallowable value in the SPCCI combustion, in the relationship between acompression ratio of the engine 1 and the SI fuel rate. As indicated ingraph 1504, in the SPCCI combustion, unless the SI fuel rate isincreased to some degree or higher degree, combustion noise cannot bereduced so as to indicate the allowable value or less, regardless of thecompression ratio of the engine. A specific value of the allowable valueof combustion noise can be determined as appropriate. In the graph 1504,the greater the compression ratio of the engine is, the higher thetemperature in the combustion chamber at the compression top dead centeris, and CI combustion may be steep. Therefore, unless the SI fuel rateis increased, the combustion noise cannot be reduced to the allowablevalue or less. When the compression ratio of the engine is small, thetemperature in the combustion chamber at the compression top dead centerbecomes low, and the CI combustion is not steep. Therefore, even whenthe SI fuel rate is small, the combustion noise can be reduced to theallowable value or less.

In this examination, as indicated in a matrix of reference numeral 1501in FIG. 15, two parameters that are the EGR ratio of the SI part, andthe total EGR ratio in the entirety of the combustion chamber arechanged, to search for a range that satisfies the above-describedconstraint conditions (1) to (3). In the illustrated example, theexternal EGR ratio is changed in increments of 5%, and the total EGRratio is changed in increments of 10%. The increments for the change ofthe EGR ratio can be set as appropriate. The searching is performed suchthat, for example, the external EGR ratio is fixed to a certain value,and, while the total EGR ratio is changed, a range of the total EGRratio that satisfies the constraint conditions (1) to (3) is searchedfor. While the value of the external EGR ratio is changed, the searchingis repeated.

The SI fuel rate in the SPCCI combustion is changed without changing therelationship between the EGR ratio of the SI part and the total EGRratio, whereby the above-described constraint conditions (1) to (3) canbe satisfied. As indicated in a matrix of reference numeral 1502, the SIfuel rate is changed with respect to the one vertical column in thematrix 1501 (that is, for each value of the external EGR ratio), therebysearching for a range of the total EGR ratio which satisfies theconstraint conditions (1) to (3).

FIG. 15 illustrates, in a graph of reference numeral 1503, a searchingresult in the matrix 1502. The graph 1503 represents an example of asearching result in the case of the compression ratio of the enginebeing 16 and the external EGR ratio being 20%. In the graph 1503 inwhich the horizontal axis represents an SI fuel rate, and the verticalaxis represents the total EGR ratio, a region (II) is a region, to theleft of an alternate long and short dash line, in which the SI fuel rateis 0.5 or less. This region corresponds to the lower limit of the SIfuel rate at which combustion noise is allowable in the SPCCI combustionwhen the compression ratio of the engine is 16, as indicated in thegraph 1504. That is, the region (II) is a region that does not satisfythe constraint condition (3). The region (II) corresponds to a region inwhich combustion noise is greater than the allowable value since theproportion of the SI combustion is small in the SPCCI combustion.

A region (III) is a region above the broken line in the graph 1503. Theregion is a region in which the total EGR ratio is great. In the region(III), the G/F of air-fuel mixture of the CI part is excessively great,whereby combustion stability cannot be assured in the CI combustion ofthe SPCCI combustion. That is, the region (III) is a region that doesnot satisfy the constraint condition (2).

A region (IV) is a region below the solid line in the graph 1503. Thisregion is a region in which the total EGR ratio is small. In the region(IV), the temperature of the CI part at the compression top dead centeris low, and autoignition of air-fuel mixture of the CI part is notstably performed in the SPCCI combustion. That is, the region (IV) isalso a region that does not satisfy the constraint condition (2).

A region (I) in the graph 1503 is a region that satisfies the constraintconditions (2) and (3).

As described above, according to the matrix 1501, the matrix 1502, andthe graph 1503, while the SI fuel rate is changed for each external EGRratio, a range of the total EGR ratio which satisfies the constraintconditions is searched for. As a result, as an example of a searchingresult of the matrix 1501, a graph of reference numeral 1505 in FIG. 15can be obtained. The graph 1505 represents a region that satisfies theconstraint conditions (1) to (3), on a plane in which the horizontalaxis represents the external EGR ratio of the SI part, and the verticalaxis represents the total EGR ratio. The graph 1503 and the graph 1505represent a range of the total EGR ratio that satisfies the constraintconditions in a certain external EGR ratio (in the illustrated example,the EGR ratio is 20%) as indicated by double headed arrows in FIG. 15.

FIG. 16 illustrates a relationship, between the G/F (horizontal axis) ofthe SI part and the total G/F (vertical axis) in the entirety of thecombustion chamber, which allows the SPCCI combustion to be stablyperformed in a state where the G/F of air-fuel mixture stratifies. FIG.16 illustrates a graph 1601 obtained when the EGR ratio in the graph1505 in FIG. 15 is replaced with the G/F for both the vertical axis andthe horizontal axis. The hatched region in the graph 1601 is a regionthat satisfies the constraint conditions. When the relationship betweenthe G/F of the SI part and the total G/F in the combustion chamber is inthis region, the SPCCI combustion is stabilized.

A line 1602 on the upper side of the region shown in the graph 1601corresponds to a line for assuring the SI fuel rate that allowscombustion noise to be avoided by serving as a limitation line abovewhich the CI part is excessively diluted and CI combustion is notstabilized. Furthermore, a line 1603 on the right side of the regioncorresponds to a limitation line for assuring stability of the SIcombustion in the SI part (that is, a line for allowing the constraintcondition (1) to be satisfied). A line 1604 on the lower side of theregion is a line for allowing a temperature of the CI part to be assuredand allowing autoignition to be stabilized. The line is a straight linethat extends in the upper right direction in the graph 1601. The lineshifts upward as indicated by the alternate long and short dash linewhen the compression ratio c of the engine is increased, and shiftsdownward as indicated by the alternate long and two short dashes linewhen the compression ratio c of the engine is reduced.

In the graph 1601, a line 1605 is further added. The line 1605 is a linefor allowing great combustion noise due to knocking to be avoided, andsatisfies G/F=18 as illustrated also in FIG. 13. The line 1605intersects the line 1604 described above. As described above, above theline 1604, normal combustion noise satisfies the allowable value in theSPCCI combustion. However, when the total EGR ratio is below the line1605, knocking (abnormal combustion) may occur, so that the line 1605needs to be prioritized.

According to the above-described examination, the G/F in the combustionchamber stratifies, whereby the total G/F for stabilizing the SPCCIcombustion is in a range from 18 to 50. At this time, the G/F of the SIpart is in a range from 14 to 22. When the G/F of the SI part is great,the total G/F needs to be great in order to stabilize the SPCCIcombustion. Furthermore, when the compression ratio c of the engine 1 isgreat, the total G/F needs to be increased as compared to when thecompression ratio c is small, in order to stabilize the SPCCIcombustion.

When the G/F stratifies, the air-fuel mixture can be further diluted ascompared to the range of the G/F which is shown in FIG. 13. Therefore,fuel economy performance of the engine is advantageously improved. TheG/F can stratify by, for example, strong swirl flow being generated inthe combustion chamber 17, the shape of the combustion chamber 17 beingproperly designed, or combination thereof being used.

In a case where the G/F in the combustion chamber stratifies, theair-fuel mixture of the SI part is set to have an excess air ratio λ of1.0±0.2 and the EGR ratio of 34% or less. As enclosed by the alternatelong and two short dashes line in FIG. 14, when air-fuel mixturesatisfies the excess air ratio λ≈1 and has the EGR ratio of 34% or less,sensitivity for change of timing of autoignition with respect to changeof ignition timing is high. That is, in a case where the G/F in thecombustion chamber stratifies, the total G/F is in a range from 18 to50, the G/F of the SI part is in a range from 14 to 22, and the excessair ratio λ≈1 in the entirety of the combustion chamber is satisfied,timing of autoignition can be accurately changed with respect to changeof ignition timing in the SPCCI combustion.

(Operation Control of Engine)

The engine 1 switches between the SI combustion and the SPCCI combustionaccording to the operation state in the operation region map 701. Theengine 1 also changes the SI rate according to the operation state ofthe engine 1. Thus, both inhibition of generation of combustion noiseand improvement of fuel economy can be achieved.

FIG. 9 illustrates change of the SI rate, change of a state quantity inthe combustion chamber 17, change of an opening period of the intakevalve and an opening period of the exhaust valve, and change of fuelinjection timing and ignition timing, according to whether the load onthe engine 1 is high or low. FIG. 9 corresponds to the operation regionmap 701 shown in FIG. 7. Hereinafter, operation control of the engine 1will be described on the assumption that load on the engine 1 isgradually enhanced at a predetermined number of revolutions.

(Low Load Region (Low Load SI Combustion))

In the low load region (A), the engine 1 performs low load SIcombustion. When the operation state of the engine 1 is in the low loadregion, the SI rate is constantly 100%.

In the low load region, as described above, the G/F of air-fuel mixtureis made constant between 18 and 50. In the engine 1, fresh air andburned gas are introduced, in amounts corresponding to an amount offuel, into the combustion chamber 17. An amount of fresh air to beintroduced is adjusted by throttling and/or Miller cycle as describedabove. Since a dilution rate is high, the temperature in the combustionchamber 17 is enhanced in order to stabilize SI combustion. In theengine 1 in the low load region, internal EGR gas is introduced into thecombustion chamber 17. An opening degree of the swirl control valve 56is adjusted as appropriate.

The internal EGR gas is introduced into the combustion chamber 17 (thatis, burned gas is confined in the combustion chamber 17) by setting anegative overlap period in which the intake valve 21 and the exhaustvalve 22 are both closed around the exhaust top dead center. The lengthof the negative overlap period is set as appropriate by an opening timeof the intake valve 21 being adjusted by the intake electric S-VT 23,and an opening time of the exhaust valve 22 being adjusted by theexhaust electric S-VT 24, whereby an amount of the internal EGR gas isadjusted. The internal EGR gas may be introduced into the combustionchamber 17 by setting a positive overlap period in which both the intakevalve 21 and the exhaust valve 22 are opened.

In the low load region, an amount for filling the combustion chamber 17is adjusted to be less than 100%. As an amount of fuel is increased, anamount of fresh air and an amount of the internal EGR gas which are tobe introduced into the combustion chamber 17 are gradually increased.The EGR ratio in the low load region is, for example, 40%.

The injector 6 injects fuel into the combustion chamber 17 in the intakestroke. In the combustion chamber 17, homogeneous air-fuel mixturehaving the excess air ratio λ of 1.0±0.2 and the G/F of 18 to 50 isformed. The excess air ratio h is preferably 1.0 to 1.2.

The spark plug 25 ignites the air-fuel mixture at predetermined timingbefore the compression top dead center. Thus, the air-fuel mixturecombusts by flame propagation without reaching autoignition. That is, inthe low load region, combustion of the air-fuel mixture is completed byonly SI combustion. As shown in FIG. 21, the ignition timing by thespark plug 25 is gradually retarded as the load on the engine 1increases.

In low load SI combustion, when the load on the engine 1 is low, theinitial temperature in the combustion chamber 17 becomes relatively lowand therefore the combustion is likely to become unstable. For thisreason, in the case of low load in the low load region, the ignitiontiming is advanced to some extent. Thus, the air-fuel mixture is ignitedat timing at which the temperature in the combustion chamber increasesby compression operation of the piston, whereby stability of the lowload SI combustion can be ensured.

On the other hand, in low load SI combustion, as the load on the engine1 increases, the initial temperature in the combustion chamber 17becomes relatively high. Therefore, although stability of the low loadSI combustion is ensured, a period from ignition by the spark plug 25until combustion occurs is shortened, and the start timing of low loadSI combustion becomes earlier. For this reason, the ignition timing isretarded as the load becomes higher in the low load region. Thus, thestart timing of low load SI combustion is delayed, whereby the starttiming can be adjusted to desired timing after the compression top deadcenter.

(Second Intermediate Load Region (Non-Supercharge SPCCI Combustion))

When load on the engine 1 becomes high and the operation state entersthe second intermediate load region (B2), the engine 1 switches from thelow load SI combustion to non-supercharge SPCCI combustion. The SI rateis less than 100%. An amount of fuel is increased according to load onthe engine 1 being increased. In the second intermediate load region,when the load is low, the proportion of CI combustion is increasedaccording to an amount of fuel being increased. The SI rate is graduallyreduced according to load on the engine 1 being increased. The SI rateis reduced to a predetermined value (lowest value) that is 50% or less,in the example shown in FIG. 9.

Since an amount of fuel is increased, a temperature of combustionbecomes high in the second intermediate load region. When thetemperature in the combustion chamber 17 is excessively high, heatgeneration at the start of CI combustion is intense. In this case,combustion noise is increased.

Therefore, in the second intermediate load region, a ratio between theinternal EGR gas and the external EGR gas is changed with respect tochange of load on the engine 1 in order to adjust a temperature (initialtemperature) in the combustion chamber 17 before start of compression.That is, according to load on the engine 1 being increased, hot internalEGR gas is gradually reduced, and cooled external EGR gas is graduallyincreased. As described above, the internal EGR gas and the external EGRgas both correspond to the operation quantity relevant to the initialtemperature in the combustion chamber 17.

The negative overlap period is changed from the maximum to zero as theload is increased in the second intermediate load region. The internalEGR gas becomes zero when the load is highest in the second intermediateload region. The same applies to a case where a positive overlap periodof the intake valve 21 and the exhaust valve 22 is set. A temperature inthe combustion chamber 17 is adjusted by the overlap period beingadjusted, resulting in the SI rate in the SPCCI combustion being able tobe adjusted.

An opening degree of the EGR valve 54 is changed in the secondintermediate load region such that the external EGR gas is increased asthe load is increased. An amount of the external EGR gas to beintroduced into the combustion chamber 17 is adjusted so as to be, forexample, 0 to 30% when represented as the EGR ratio. In the secondintermediate load region, the EGR gas is replaced such that the internalEGR gas is replaced with the external EGR gas according to load on theengine 1 being increased. Also by adjusting the EGR ratio, thetemperature in the combustion chamber 17 is adjusted. Therefore, the SIrate in the SPCCI combustion can be adjusted.

An amount of EGR gas to be introduced into the combustion chamber 17 issequential between the low load region and the second intermediate loadregion. In a region, of the second intermediate load region, in whichload is low, similarly to the low load region, a large amount ofinternal EGR gas is introduced into the combustion chamber 17. Since thetemperature in the combustion chamber 17 becomes high, when load on theengine 1 is low, autoignition of air-fuel mixture assuredly occurs. In aregion, of the second intermediate load region, in which load is high, alarge amount of external EGR gas is introduced into the combustionchamber 17. Since the initial temperature in the combustion chamber 17becomes low, when load on the engine 1 is high, combustion noise causedby CI combustion can be reduced.

In the second intermediate load region, an amount for filling thecombustion chamber 17 is 100%. The opening degree of the throttle valve43 is a fully opened degree. An amount of EGR gas which is the sum ofinternal EGR gas and external EGR gas is adjusted, whereby an amount offresh air to be introduced into the combustion chamber 17 is adjusted soas to correspond to an amount of fuel.

In non-supercharge SPCCI combustion, as the proportion of the CIcombustion is increased, timing of autoignition is advanced. If thetiming of autoignition is earlier than the compression top dead center,heat generation at the start of CI combustion is intense. In this case,combustion noise is increased. Therefore, when load on the engine 1reaches a predetermined load L1, the engine 1 gradually increases the SIrate according to load on the engine 1 being increased.

That is, as an amount of fuel is increased, the engine 1 increases theproportion of SI combustion. Specifically, as shown in the upper diagramof FIG. 10, in non-supercharge SPCCI combustion, as the load on theengine 1 increases (as the amount of fuel increases), the ignitiontiming is gradually advanced.

As described above, an amount of internal EGR gas to be introduced isreduced and an amount of external EGR gas to be introduced is increased,whereby the initial temperature in the combustion chamber 17 isadjusted. Therefore, even if the SI rate is increased according to anamount of fuel being increased, enhancement in temperature at thecompression top dead center can be inhibited. In a region in which theload is the predetermined load L1 or higher, as shown in FIG. 22, as theload increases, the initial temperature in the combustion chamber 17decreases. From the above, the slope of the heat generation rate in SIcombustion hardly changes even if the load on the engine 1 increases.Therefore, when the ignition timing is advanced, the amount of heatgeneration in the SI combustion is increased according to start of theSI combustion becoming earlier, resulting in increase in the SI rate.The SI rate increases as the initial temperature in the combustionchamber 17 decreases.

As a result of enhancement in temperature in the combustion chamber 17by the SI combustion being inhibited, autoignition of unburned air-fuelmixture occurs at timing at or after the compression top dead center.Since an amount of heat generated by the SI combustion is increased,heat generation in the CI combustion is almost the same even if load onthe engine 1 is increased. Therefore, the SI rate is set to be graduallyincreased according to load on the engine 1 being increased, wherebyincrease of combustion noise can be avoided. The higher load is, themore greatly the center of gravity of combustion in non-superchargeSPCCI combustion is retarded.

In the second intermediate load region, the swirl control valve 56 isfully closed or opened to a predetermined opening degree on the closingside. In the combustion chamber 17, strong swirl flow is formed at aswirl ratio of 4 or greater. Thus, residual gas in the cavity 31 isexpelled to the outside of the cavity 31.

In the second intermediate load region, the injector 6 injects fuel intothe combustion chamber 17 in two parts as preceding injection andsucceeding injection in the compression stroke. In the precedinginjection, fuel is injected at timing that is not close to the ignitiontiming. In the succeeding injection, fuel is injected at timing close tothe ignition timing. When the injector 6 performs the precedinginjection, since the piston 3 is distant from the top dead center, theinjected fuel spray reaches the outside of the cavity 31 on the topsurface of the piston 3 that is being moved toward the top dead centerin the upward direction. A squish area 171 is formed in a region outsidethe cavity 31 (see FIG. 2). Fuel injected in the preceding injection isaccumulated in the squish area 171 while the piston 3 is moving upward,to form air-fuel mixture in the squish area 171.

When the injector 6 performs the succeeding injection, since the piston3 is positioned close to the top dead center, the injected fuel sprayenters the cavity 31. The fuel injected in the succeeding injectionforms air-fuel mixture in a region inside the cavity 31. The “regioninside the cavity 31” may be a region obtained by addition of a regioninside the cavity 31, and a region up to the opening of the cavity 31from a projected surface obtained by the opening of the cavity 31 beingprojected on the roof of the combustion chamber 17. The region insidethe cavity 31 may be referred to as a region, other than the squish area171, in the combustion chamber 17. Fuel is almost uniformly distributedover the entirety of the combustion chamber 17.

Gas flows in the region inside the cavity 31 according to fuel beinginjected into the cavity 31 in succeeding injection. The turbulentenergy in the combustion chamber 17 is attenuated according to progressof the compression stroke when a time up to the ignition timing is long.However, the injection timing in the succeeding injection is closer tothe ignition timing than the injection timing in the precedinginjection. Therefore, the spark plug 25 can ignite air-fuel mixture inthe region inside the cavity 31 in a state where the turbulent energy inthe cavity 31 remains high. Thus, a combustion speed in the SIcombustion is enhanced. When the combustion speed in the SI combustionis enhanced, the SI combustion is stabilized, whereby controllability ofthe CI combustion by the SI combustion is enhanced.

In the entirety of the combustion chamber 17, air-fuel mixture has theexcess air ratio λ of 1.0±0.2 and the G/F of 18 to 50. The residual gashas been expelled from the cavity 31, whereby the G/F of air-fuelmixture near the spark plug 25 is 14 to 22. The G/F in the combustionchamber 17 stratifies. Meanwhile, since fuel is almost uniformlydistributed, improvement of fuel economy by reduction of unburned fuelloss, and improvement of exhaust gas performance by avoiding generationof smoke, can be achieved. In the entirety of the combustion chamber 17,the excess air ratio λ is preferably 1.0 to 1.2.

At predetermined timing before the compression top dead center, thespark plug 25 ignites air-fuel mixture, whereby the air-fuel mixture iscombusted by flame propagation. Thereafter, autoignition of the unburnedair-fuel mixture occurs at target timing, to cause CI combustion. Thefuel injected in the succeeding injection is subjected mainly to SIcombustion. The fuel injected in the preceding injection is subjectedmainly to CI combustion. Since the preceding injection is performed inthe compression stroke, abnormal combustion such as preignition can beprevented from being caused by the fuel injected in the precedinginjection. Furthermore, fuel injected in the succeeding injection can bestably combusted by flame propagation. The G/F of air-fuel mixture inthe combustion chamber 17 stratifies, and the G/F in the entirety of thecombustion chamber 17 is set to 18 to 50, whereby the SPCCI combustioncan be stably performed.

(First Intermediate Load Region (Supercharge SPCCI Combustion))

When load on the engine 1 is further increased and the operation stateof the engine 1 thus enters the first intermediate load region (B1), thesupercharger 44 performs supercharging with fresh air and external EGRgas. An amount of fresh air and an amount of external EGR gas, which areto be introduced into the combustion chamber 17, are both increasedaccording to load on the engine 1 being increased. An amount of theexternal EGR gas to be introduced into the combustion chamber 17 is, forexample, 30% when represented as the EGR ratio. The EGR ratio is almostconstant regardless of whether load on the engine 1 is high or low.Therefore, the G/F of air-fuel mixture is also almost constantregardless of whether load on the engine 1 is high or low. An amount ofEGR gas to be introduced into the combustion chamber 17 is sequentialbetween the second intermediate load region and the first intermediateload region. In the first intermediate load region, the initialtemperature in the combustion chamber 17 is lower than in the secondintermediate load region.

The SI rate indicates a predetermined value that is less than 100%, andis made constant or almost constant regardless of whether the load onthe engine 1 is high or low. When the SI rate in the second intermediateload region, particularly, the SI rate for which load is higher than thepredetermined load L1, and which is gradually increased according toload on the engine 1 being increased, is compared with the SI rate inthe first intermediate load region, the SI rate in the firstintermediate load region in which load on the engine 1 is higher, isgreater than the SI rate in the second intermediate load region. The SIrate is sequential at a boundary between the first intermediate loadregion and the second intermediate load region.

In the first intermediate load region, the SI rate may be slightlychanged according to load on the engine 1 being changed. A change rateof change of the SI rate to change of load on the engine 1 in the firstintermediate load region may be set to be less than a change rate of theSI rate on the high load side of the second intermediate load region.

As shown in the lower diagram of FIG. 10, and FIG. 21, also insupercharge SPCCI combustion, the ignition timing is gradually advancedaccording to an amount of fuel being increased. As described above, anamount of fresh air and an amount of EGR gas, which are to be introducedinto the combustion chamber 17, are increased by supercharging, wherebyheat capacity is high. Even when an amount of fuel is increased,enhancement in temperature in the combustion chamber due to SIcombustion can be inhibited. A waveform of the heat generation rate insupercharge SPCCI combustion is enlarged in similar shapes according toload being increased.

That is, an amount of heat generated by SI combustion is increased withlittle change of the slope of the heat generation rate in the SIcombustion. At almost the same timing at or after the compression topdead center, autoignition of unburned air-fuel mixture occurs. An amountof heat generated by CI combustion is increased when load on the engine1 becomes high. As a result, in the first intermediate load region, bothan amount of heat generated by the SI combustion and an amount of heatgenerated by the CI combustion are increased, whereby the SI rate isconstant regardless of whether load on the engine 1 is high or low. Whena peak of heat generation in the CI combustion is enhanced, combustionnoise is increased. However, in the first intermediate load region,since load on the engine 1 is relatively high, combustion noise can beallowed to a certain degree. The higher load is, the more greatly thecenter of gravity of combustion in supercharge SPCCI combustion isretarded.

In the first intermediate load region, a positive overlap period inwhich both the intake valve 21 and the exhaust valve 22 are opened isset around the exhaust top dead center. Burned gas in the combustionchamber 17 is scavenged by supercharging. Thus, since the temperature inthe combustion chamber 17 becomes low, abnormal combustion can beinhibited from occurring when load on the engine 1 is relatively high.Furthermore, by the temperature in the combustion chamber 17 beinglowered, timing of autoignition can be made appropriate in a region inwhich load on the engine 1 is relatively high, and the SI rate can bemaintained as a predetermined SI rate. That is, by the overlap periodbeing adjusted, the SI rate can be adjusted. Furthermore, by scavengingburned gas, an amount of fresh air to be filled in the combustionchamber 17 can be increased.

In the first intermediate load region, the swirl control valve 56 isfully closed or opened to a predetermined opening degree on the closingside. In the combustion chamber 17, strong swirl flow is formed at aswirl ratio of 4 or greater. Thus, residual gas in the cavity 31 isexpelled to the outside of the cavity 31.

In the first intermediate load region, similarly to the secondintermediate load region, the injector 6 injects fuel into thecombustion chamber 17 in two parts as preceding injection and succeedinginjection in the compression stroke. In the preceding injection, fuel isinjected at timing which is not close to the ignition timing. In thesucceeding injection, fuel is injected at timing close to the ignitiontiming. The fuel is almost uniformly distributed over the entirety ofthe combustion chamber 17. In the entirety of the combustion chamber 17,air-fuel mixture has the excess air ratio λ of 1.0±0.2 and the G/F of 18to 50. Since residual gas has been expelled from the cavity 31, the G/Fof air-fuel mixture near the spark plug 25 is 14 to 22. The G/F in thecombustion chamber 17 stratifies. Meanwhile, since fuel is almostuniformly distributed over the entirety of the combustion chamber 17,improvement of fuel economy by reduction of unburned fuel loss, andimprovement of exhaust gas performance by avoiding generation of smoke,can be achieved. In the entirety of the combustion chamber 17, theexcess air ratio h is preferably 1.0 to 1.2.

At predetermined timing before the compression top dead center, thespark plug 25 ignites air-fuel mixture, whereby the air-fuel mixture iscombusted by flame propagation. Thereafter, autoignition of unburnedair-fuel mixture occurs at target timing, to cause CI combustion. Fuelinjected in the succeeding injection is subjected mainly to SIcombustion. Fuel injected in the preceding injection is subjected mainlyto CI combustion. Since the preceding injection is performed in thecompression stroke, abnormal combustion such as preignition can beprevented from being caused by fuel injected in the preceding injection.Furthermore, fuel injected in the succeeding injection can be stablycombusted by flame propagation. When the G/F of air-fuel mixture in thecombustion chamber 17 stratifies, and the G/F in the entirety of thecombustion chamber 17 is set to 18 to 50, the SPCCI combustion can bestably performed

When load on the engine 1 is further increased and the operation stateof the engine 1 enters the high load region (C), the engine 1 performshigh load SI combustion. Therefore, the SI rate in the high load regionis 100%.

The throttle valve 43 is fully opened. The supercharger 44 performssupercharging with fresh air and external EGR gas also in the high loadregion. The opening degree of the EGR valve 54 is adjusted, whereby anamount of external EGR gas to be introduced is gradually reducedaccording to load on the engine 1 being increased. Thus, fresh air to beintroduced into the combustion chamber 17 is increased when load on theengine 1 becomes high. When an amount of fresh air is increased, anamount of fuel can be increased. Therefore, the maximum output of theengine 1 is advantageously enhanced. An amount of EGR gas to beintroduced into the combustion chamber 17 is sequential between thefirst intermediate load region and the high load region.

Also in the high load region, similarly to the first intermediate loadregion, a positive overlap period in which both the intake valve 21 andthe exhaust valve 22 are opened is set around the exhaust top deadcenter. Burned gas in the combustion chamber 17 is scavenged bysupercharging. Thus, abnormal combustion is inhibited. Furthermore, anamount of fresh air to be filled in the combustion chamber 17 can beincreased.

In a region (that is, the first high load region (C1)) on the lowrotation side of the high load region, the injector 6 injects fuel intothe combustion chamber 17 in the retard period as described above. In aregion (that is, the second high load region (C2)) on the high rotationside of the high load region, the injector 6 injects fuel into thecombustion chamber 17 in the intake stroke. In either case, almosthomogeneous air-fuel mixture having the excess air ratio λ of 1.0±0.2and the G/F of 18 to 50 is formed in the combustion chamber 17. At thehighest load, the excess air ratio λ may be, for example, 0.8.Furthermore, the G/F of air-fuel mixture may be, for example, 17 at thehighest load. At predetermined timing before the compression top deadcenter, the spark plug 25 ignites air-fuel mixture, whereby the air-fuelmixture is combusted by flame propagation. In the high load region, SIcombustion of the air-fuel mixture occurs by high-pressure retardinjection or fuel injection in the intake stroke without causingautoignition.

The spark plug 25 ignites air-fuel mixture at predetermined timing nearthe compression top dead center. Thus, the air-fuel mixture combusts byflame propagation. In the high load region, the air-fuel mixtureundergoes SI combustion without reaching autoignition, by high-pressureretard injection, or fuel injection in the intake stroke. That is,combustion of the air-fuel mixture is completed by only SI combustion.

The ignition timing is retarded from the ignition timing in the firstintermediate load region, and is set to be after the fuel injectiontiming within a period from the later stage of the compression stroke tothe initial stage of the expansion stroke. In a high load SI region, asshown in FIG. 21, as the load on the engine 1 increases, the ignitiontiming is gradually retarded. Thus, the combustion speed of the air-fuelmixture becomes fast in accordance with the increase in the load on theengine 1, so that start of SI combustion is delayed accordingly, wherebyoccurrence of abnormal combustion such as preignition or knocking can beavoided.

(Adjustment of SI Rate)

FIG. 11 shows a flow of operation control, of the engine, performed bythe ECU 10. The ECU 10 determines an operation state of the engine 1based on the detection signals from the sensors SW1 to SW16, and adjustsa state quantity in the combustion chamber 17, adjusts an injectionamount, adjusts injection timing, and adjusts ignition timing such thatcombustion in the combustion chamber 17 becomes combustion at an SI ratecorresponding to the operation state. The ECU 10 also adjusts the SIrate when determining that the SI rate needs to be adjusted, based ondetection signals from the sensors.

The ECU firstly reads detection signals from the sensors SW1 to SW16 instep S1. Subsequently, the ECU 10 determines an operation state of theengine 1 based on the detection signals, and sets a target SI rate instep S2. The target SI rate is as shown in FIG. 9.

In subsequent step S3, the ECU 10 sets a target in-cylinder statequantity for obtaining the target SI rate having been set, based on apreset combustion model. Specifically, a target temperature, a targetpressure, and a target state quantity in the combustion chamber 17 areset. In step S4, the ECU 10 sets an opening degree of the EGR valve 54,an opening degree of the throttle valve 43, an opening degree of the airbypass valve 48, an opening degree of the swirl control valve 56, andphase angles of the intake electric S-VT 23 and the exhaust electricS-VT 24, which are necessary for obtaining the target in-cylinder statequantity. The ECU 10 presets control amounts for these devices, and setsthe control amounts based on a map stored in the ECU 10. The ECU 10outputs control signals to the EGR valve 54, the throttle valve 43, theair bypass valve 48, the swirl control valve 56, and the intake electricS-VT 23 and the exhaust electric S-VT 24, based on the set controlamounts. Each device operates based on the control signal from the ECU10, whereby the state quantity in the combustion chamber 17 becomes thetarget state quantity.

The ECU 10 further calculates a predicted value and an estimated valueof the state quantity in the combustion chamber 17, based on the setcontrol amount for each device. The predicted value of the statequantity is a value obtained by predicting a state quantity, in thecombustion chamber 17, before closing of the intake valve 21, and isused for setting an amount of fuel injected in the intake stroke asdescribed below. The estimated value of the state quantity is a valueobtained by estimating a state quantity, in the combustion chamber 17,after closing of the intake valve 21, and is used for setting an amountof fuel injected in the compression stroke and setting ignition timingas described below. The estimated value of the state quantity is alsoused for calculating an error in state quantity based on comparison withan actual combustion state as described below.

The ECU 10 sets, in step S5, an amount of fuel to be injected in theintake stroke based on the predicted value of the state quantity. Whenfuel is not injected in the intake stroke, the amount of injected fuelis zero. In step S6, the ECU 10 controls injection by the injector 6.That is, a control signal is outputted to the injector 6 such that fuelis injected into the combustion chamber 17 at predetermined injectiontiming.

In step S7, the ECU 10 sets an amount of fuel to be injected in thecompression stroke based on the estimated value of the state quantityand a result of injection of fuel in the intake stroke. When fuel is notinjected in the compression stroke, the amount of injected fuel is zero.When divided injection is performed in the compression stroke, each ofan amount of injection in the preceding injection and an amount ofinjection in the succeeding injection is set. In step S8, the ECU 10outputs a control signal to the injector 6 such that fuel is injectedinto the combustion chamber 17 at injection timing based on the presetmap.

In step S9, the ECU 10 sets ignition timing based on the estimated valueof the state quantity and a result of injection of fuel in thecompression stroke. In step S10, the ECU 10 outputs a control signal tothe spark plug 25 such that air-fuel mixture in the combustion chamber17 is ignited at the set ignition timing.

The spark plug 25 ignites the air-fuel mixture, to cause SI combustionor SPCCI combustion in the combustion chamber 17. In step S11, the ECU10 reads change of pressure, in the combustion chamber 17, detected bythe pressure sensor SW6, and determines, based thereon, a combustionstate of the air-fuel mixture in the combustion chamber 17. Furthermore,in step S12, the ECU 10 compares a result of detection of the combustionstate with the estimated value of the state quantity, which is estimatedin step S4, and calculates an error between the estimated value of thestate quantity and an actual state quantity. The calculated error isused for estimation in step S4 in cycles subsequent to this cycle. TheECU 10 adjusts opening degrees of the throttle valve 43, the EGR valve54, the swirl control valve 56, and/or the air bypass valve 48, andphase angles of the intake electric S-VT 23 and the exhaust electricS-VT 24 so as to remove the error in the state quantity. Thus, amountsof fresh air and EGR gas to be introduced into the combustion chamber 17are adjusted. Feedback of the error in the state quantity corresponds toadjustment of an SI rate performed by the ECU 10 when adjustment of theSI rate is determined to be necessary based on the error between thetarget SI rate and the actual SI rate.

Furthermore, in step S8, injection timing in the compression stroke isadvanced so as to be earlier than injection timing based on the map, bythe ECU 10, such that ignition timing can be advanced, when the ECU 10makes prediction that the temperature in the combustion chamber 17 islower than the target temperature based on the estimated value of thestate quantity. Meanwhile, in step S8, the ECU 10 retards injectiontiming in the compression stroke from injection timing based on the mapsuch that ignition timing can be retarded when the ECU 10 makesprediction that the temperature in the combustion chamber 17 is higherthan the target temperature based on the estimated value of the statequantity.

That is, as indicated by P2 in FIG. 12, when the temperature in thecombustion chamber 17 is low, after SI combustion by spark ignitionstarts, timing θ_(CI) at which autoignition of unburned air-fuel mixtureoccurs delays, and the SI rate is deviated from the target SI rate (seeP1). In this case, unburned fuel is increased or exhaust gas performanceis degraded.

Therefore, when the ECU 10 makes prediction that the temperature in thecombustion chamber 17 is lower than the target temperature, the ECU 10advances the injection timing and also advances the ignition timingθ_(IG) in step S10 shown in FIG. 11. As indicated by P3 in FIG. 12, thestart of SI combustion becomes earlier, whereby sufficient heatgeneration by the SI combustion can be caused. Therefore, when thetemperature in the combustion chamber 17 is low, the timing θ_(CI) ofautoignition of unburned air-fuel mixture can be prevented from beingdelayed. As a result, the SI rate approaches the target SI rate.Increase of unburned fuel or degradation of exhaust gas performance areprevented.

Furthermore, as indicated by P4 in FIG. 12, when the temperature in thecombustion chamber 17 is high, autoignition of unburned air-fuel mixtureoccurs immediately after the SI combustion by spark ignition starts, andthe SI rate is deviated from the target SI rate (see P1). In this case,combustion noise is increased.

Therefore, when the ECU 10 makes prediction that the temperature in thecombustion chamber 17 is higher than the target temperature, the ECU 10retards the injection timing and also retards the ignition timing θ_(IG)in step S10 shown in FIG. 11. As indicated by P5 in FIG. 12, since startof the SI combustion is delayed, when the temperature in the combustionchamber 17 is high, the timing θ_(CI) of autoignition of unburnedair-fuel mixture can be prevented from being earlier. As a result, theSI rate approaches the target SI rate. Increase of combustion noise isprevented.

These adjustments of the injection timing and the ignition timingcorrespond to adjustment of the SI rate performed by the ECU 10 whenadjustment of the SI rate in the SPCCI combustion is determined to benecessary. By adjusting the injection timing, an appropriate air-fuelmixture can be formed in the combustion chamber 17 at the advanced orretarded ignition timing. The spark plug 25 can assuredly ignite theair-fuel mixture, and autoignition of unburned air-fuel mixture can bealso performed at appropriate timing.

In FIG. 12, adjustment of the state quantity in the combustion chamber17 by controlling the throttle valve 43, the EGR valve 54, the airbypass valve 48, the swirl control valve 56, the intake electric S-VT23, and the exhaust electric S-VT 24 based on an actual combustion stateis as described for step S12 and step S4 in FIG. 11.

The engine 1 adjusts the SI rate by means of the state quantity settingdevice that includes the throttle valve 43, the EGR valve 54, the airbypass valve 48, the swirl control valve 56, the intake electric S-VT23, and the exhaust electric S-VT 24. By adjusting the state quantity inthe combustion chamber 17, the SI rate can be roughly adjusted. Inaddition thereto, the engine 1 adjusts the fuel injection timing and theignition timing, to adjust the SI rate. By adjusting the injectiontiming and the ignition timing, for example, correction of differencebetween the cylinders can be performed, and slight adjustment of theautoignition timing can be performed. By adjusting the SI rate at twostages, the engine 1 can accurately perform the target SPCCI combustioncorresponding to the operation state.

The control of the engine 1 by the ECU 10 is not limited to controlbased on the combustion model described above.

(Another Exemplary Configuration of Operation Region Map for Engine)

FIG. 7 illustrates, in the lower diagram, another exemplaryconfiguration of an operation region map for the engine 1. An operationregion map 702 for the engine 1 is divided into five regions accordingto whether load is high or low and whether the number of revolutions isgreat or small. Specifically, the five regions are: a low load region(1)-1 which includes an idling operation region and extends in a lowrotation region and an intermediate rotation region; an intermediateload region (1)-2 in which load is higher than that in the low loadregion, and which extends in the low rotation region and theintermediate rotation region; an intermediate rotation region (2), inthe high load region, in which load is higher than that in theintermediate load region (1)-2, and which includes full load; a lowrotation region (3), in the high load region, in which the number ofrevolutions is less than that in the intermediate rotation region (2);and a high rotation region (4) in which the number of revolutions isgreater than those in the low load region (1)-1, the intermediate loadregion (1)-2, the high-load intermediate-rotation region (2), and thehigh-load low-rotation region (3). The low rotation region, theintermediate rotation region, and the high rotation region may be a lowrotation region, an intermediate rotation region, and a high rotationregion obtained by the total operation region of the engine 1 beingdivided in the number-of-revolution direction into almost three equalportions that are the low rotation region, the intermediate rotationregion, and the high rotation region. In the example shown in FIG. 7,the low rotation represents a rotation in which the number ofrevolutions is less than N1, the high rotation represents a rotation inwhich the number of revolutions is N2 or greater, and the intermediaterotation represents a rotation in which the number of revolutions is Nior greater and less than N2. The number of revolutions Ni may be, forexample, about 1200 rpm, and the number of revolutions N2 may be, forexample, about 4000 rpm.

The engine 1 performs combustion by compression autoignition in the lowload region (1)-1, the intermediate load region (1)-2, and the high-loadintermediate-rotation region (2) in the operation region map 702, mainlyin order to improve fuel economy and improve exhaust gas performance.The operation region map 702 is different from the operation region map701 in that, in the operation region map 702, when the engine 1 operateswith a low load, and also when the engine 1 operates with a high load,the engine 1 performs the SPCCI combustion. Furthermore, the engine 1performs combustion by spark ignition in other regions, specifically, inthe high-load low-rotation region (3) and the high rotation region (4).The operation of the engine 1 in each region will be described below indetail according to a fuel injection time and an ignition time shown inFIG. 17.

(Low Load Region (1)-1)

When the engine 1 operates in the low load region (1)-1, the engine 1performs the SPCCI combustion.

In FIG. 17, reference numeral 601 represents examples of a fuelinjection time (reference numeral 6011, 6012) and an ignition time(reference numeral 6013), and a combustion waveform (that is, waveformrepresenting change of a heat generation rate with respect to a crankangle, reference numeral 6014) in the case of the engine 1 operating inan operation state of reference numeral 601 in the low load region(1)-1.

When the engine 1 operates in the low load region (1)-1, strong swirlflow is formed in the combustion chamber 17. When the engine 1 operatesin the low load region (1)-1, the swirl ratio is 4 or greater. The swirlflow is strong in the outer circumferential portion of the combustionchamber 17, and weak in the center portion. The swirl control valve(SCV) 56 is fully closed or opened to a predetermined opening degree onthe closing side. As described above, the intake port 18 is a tumbleport. Therefore, diagonal swirl flow including a tumble component and aswirl component is formed in the combustion chamber 17.

When the engine 1 operates in the low load region (1)-1, an air-fuelratio (A/F) of air-fuel mixture is leaner than the theoretical air-fuelratio in the entirety of the combustion chamber 17. That is, in theentirety of the combustion chamber 17, the excess air ratio λ of theair-fuel mixture is greater than 1. More specifically, in the entiretyof the combustion chamber 17, the A/F of the air-fuel mixture is 30 orgreater. Thus, generation of RawNOx can be inhibited, and exhaust gasperformance can be improved.

The EGR system 55 introduces EGR gas into the combustion chamber 17 asappropriate when the engine 1 operates in the low load region (1)-1.

When the engine 1 operates in the low load region (1)-1, air-fuelmixture stratifies between the center portion and the outercircumferential portion in the combustion chamber 17. The center portionof the combustion chamber 17 is a portion at which the spark plug 25 isdisposed. The outer circumferential portion of the combustion chamber 17is around the center portion, and is in contact with a liner of thecylinder 11. The center portion of the combustion chamber 17 may bedefined as a portion in which swirl flow is weak, and the outercircumferential portion thereof may be defined as a portion in whichswirl flow is strong.

The concentration of fuel in the air-fuel mixture in the center portionis higher than the concentration of fuel in the outer circumferentialportion. Specifically, the A/F of the air-fuel mixture in the centerportion is in a range from 20 to 30, and the A/F of the air-fuel mixturein the outer circumferential portion is 35 or greater.

When the engine 1 operates in the low load region (1)-1, the injector 6basically injects fuel into the combustion chamber 17 in a plurality ofparts in the compression stroke. Air-fuel mixture stratifies in thecenter portion and the outer circumferential portion of the combustionchamber 17 due to the divided injection of the fuel and strong swirlflow in the combustion chamber 17.

After the fuel injection has been ended, the spark plug 25 ignites theair-fuel mixture in the center portion of the combustion chamber 17 atpredetermined timing before the compression top dead center (seereference numeral 6013). Since the air-fuel mixture in the centerportion has a relatively high fuel concentration, ignitability isimproved, and SI combustion by flame propagation is also stabilized.When the SI combustion is stabilized, CI combustion starts atappropriate timing. In the SPCCI combustion, controllability of the CIcombustion is improved. As a result, when the engine 1 operates in thelow load region (1)-1, both inhibition of generation of combustion noiseand improvement of fuel economy performance due to a combustion periodbeing shortened can be achieved.

(Intermediate Load Region (1)-2)

Also when the engine 1 operates in the intermediate load region (1)-2,similarly to the low load region (1)-1, the engine 1 performs the SPCCIcombustion. The intermediate load region (1)-2 corresponds to theintermediate load region (B) in the operation region map 701.

In FIG. 17, reference numeral 602 represents examples of a fuelinjection time (reference numeral 6021, 6022) and an ignition time(reference numeral 6023), and a combustion waveform (reference numeral6024) in the case of the engine 1 operating in an operation state ofreference numeral 602 in the intermediate load region (1)-2.

The EGR system 55 introduces EGR gas into the combustion chamber 17 whenan operation state of the engine 1 is in the intermediate load region(1)-2.

Also when the engine 1 operates in the intermediate load region (1)-2,similarly to the low load region (1)-1, strong swirl flow is formed at aswirl ratio of 4 or greater in the combustion chamber 17. The swirlcontrol valve (SCV) 56 is fully closed or opened to a predeterminedopening degree on the closing side. When swirl flow is made strong,residual gas in the cavity 31 can be expelled from the cavity 31. As aresult, the G/F of air-fuel mixture of the SI part near the spark plug25 and the G/F of air-fuel mixture of the CI part around the SI part canbe made different from each other. Thus, as described above, when thetotal G/F in the entirety of the combustion chamber 17 is in a rangefrom 18 to 50, the SPCCI combustion can be stabilized.

Furthermore, when swirl flow is made strong, turbulent energy in thecombustion chamber 17 is enhanced. Therefore, when the engine 1 operatesin the intermediate load region (1)-2, flame propagates quickly in SIcombustion, to stabilize the SI combustion. When the SI combustion isstabilized, controllability of CI combustion is enhanced. When timing ofthe CI combustion in the SPCCI combustion is made appropriate,generation of combustion noise can be inhibited, and fuel economyperformance can be also improved. Furthermore, variation in torquebetween cycles can be reduced.

When the engine 1 operates in the intermediate load region (1)-2, theair-fuel ratio (A/F) of air-fuel mixture is the theoretical air-fuelratio (A/F≈14.7) in the entirety of the combustion chamber 17. Exhaustgas discharged from the combustion chamber 17 is purified by a three-waycatalyst, whereby exhaust gas performance of the engine 1 becomes good.The A/F of the air-fuel mixture may be set so as to fall within apurification window of the three-way catalyst. Therefore, the excess airratio K of the air-fuel mixture may be 1.0±0.2.

When the engine 1 operates in the intermediate load region (1)-2, theinjector 6 performs fuel injection (reference numeral 6021) in theintake stroke and fuel injection (reference numeral 6022) in thecompression stroke. By a first injection 6021 in the intake stroke, fuelcan be almost uniformly distributed in the combustion chamber 17. By asecond injection 6022 in the compression stroke, a temperature in thecombustion chamber 17 can be lowered due to latent heat of vaporizationof fuel. Preignition of the air-fuel mixture that contains fuel injectedin the first injection 6021 can be prevented.

When the injector 6 performs the first injection 6021 in the intakestroke and the second injection 6022 in the compression stroke, theair-fuel mixture having the excess air ratio λ of 1.0±0.2 is formed inthe combustion chamber 17 over the entirety. Since the concentration offuel in the air-fuel mixture is almost uniform, improvement of fueleconomy by reducing of unburned fuel loss and improvement of exhaust gasperformance by avoiding generation of smoke, can be achieved. The excessair ratio λ is preferably 1.0 to 1.2. Furthermore, the total G/F in theentirety of the combustion chamber 17 is in a range from 18 to 50, andthe G/F of the SI part near the spark plug 25 is 14 to 22.

At predetermined timing before the compression top dead center, thespark plug 25 ignites air-fuel mixture (reference numeral 6023), wherebythe air-fuel mixture is combusted by flame propagation. After start ofthe combustion by flame propagation, autoignition of unburned air-fuelmixture occurs at target timing, to cause CI combustion. Fuel injectedin the succeeding injection is subjected mainly to SI combustion. Fuelinjected in the preceding injection is subjected mainly to CIcombustion. The total G/F of the entirety of the combustion chamber 17is in a range from 18 to 50, and the G/F of the SI part near the sparkplug 25 is 14 to 22, whereby the SPCCI combustion can be stabilized.

As shown in the lower diagram in FIG. 7, a region in which thesupercharger 44 is off (see S/C OFF) is a part of the low load region(1)-1 and a part of the intermediate load region (1)-2. Specifically, ina region on the low rotation side in the low load region (1)-1, thesupercharger 44 is off. In a region on the high rotation side in the lowload region (1)-1, the supercharger 44 is on and boost pressure isenhanced in order to assure an amount of intake air, to be filled, whichis necessary according to the number of revolutions of the engine 1being increased. Furthermore, in a region on the low-load low-rotationside in the intermediate load region (1)-2, the supercharger 44 is off.In a region on the high load side in the intermediate load region (1)-2,the supercharger 44 is on in order to assure an amount of intake air, tobe filled, which is necessary according to an amount of injected fuelbeing increased. In the region on the high rotation side, thesupercharger 44 is on in order to assure an amount of intake air, to befilled, which is necessary according to the number of revolutions of theengine 1 being increased.

In each of the high-load intermediate-rotation region (2), the high-loadlow-rotation region (3), and the high rotation region (4), thesupercharger 44 is on over the entirety of the region

(High-Load Intermediate-Rotation Region (2))

Also when the engine 1 operates in the high-load intermediate-rotationregion (2), similarly to the low load region (1)-1 and the intermediateload region (1)-2, the engine 1 performs the SPCCI combustion.

In FIG. 17, reference numeral 603 represents examples of a fuelinjection time (reference numeral 6031, 6032) and an ignition time(reference numeral 6033), and a combustion waveform (reference numeral6034) in the case of the engine 1 operating in an operation state ofreference numeral 603 in the high-load intermediate-rotation region (2).Furthermore, in FIG. 17, reference numeral 604 represents examples of afuel injection time (reference numeral 6041) and an ignition time(reference numeral 6042), and a combustion waveform (reference numeral6043) in the case of the number of revolutions being higher than that inthe operation state of reference numeral 603.

The EGR system 55 introduces EGR gas into the combustion chamber 17 whenthe operation state of the engine 1 is in the high-loadintermediate-rotation region (2). In the engine 1, an amount of EGR gasis reduced according to load being increased. At full load, the EGR gasmay be zero.

Furthermore, also when the engine 1 operates in the high-loadintermediate-rotation region (2), similarly to the low load region(1)-1, strong swirl flow is formed at a swirl ratio of 4 or greater inthe combustion chamber 17. The swirl control valve (SCV) 56 is fullyclosed or opened to a predetermined opening degree on the closing side.

When the engine 1 operates in the high-load intermediate-rotation region(2), the air-fuel ratio (A/F) of air-fuel mixture is the theoreticalair-fuel ratio or richer than the theoretical air-fuel ratio in theentirety of the combustion chamber 17 (that is, the excess air ratio λof the air-fuel mixture satisfies λ≤1).

When the engine 1 operates on the low rotation side in the high-loadintermediate-rotation region (2), the injector 6 injects fuel in theintake stroke (reference numeral 6031) and also injects fuel at theterminal stage of the compression stroke (reference numeral 6032). Theterminal stage of the compression stroke may be a terminal stageobtained when the compression stroke is divided into three equal stagesthat are an initial stage, an intermediate stage, and the terminalstage.

In the preceding injection 6031 that starts in the intake stroke, fuelinjection may be started in the former half period of the intake stroke.The former half period of the intake stroke may be a former half periodobtained when the intake stroke is divided into two equal periods thatare the former half period and a latter half period. Specifically, inthe preceding injection, fuel injection may be started at 280° CA beforethe top dead center.

When injection starts in the preceding injection 6031 in the former halfperiod of the intake stroke, fuel spray is applied to an opening edgeportion of the cavity 31, and a part of fuel enters the squish area 171of the combustion chamber 17 and the remaining fuel enters the regioninside the cavity 31, which is not shown. The swirl flow is strong inthe outer circumferential portion of the combustion chamber 17, and isweak in the center portion. Therefore, the part of fuel that has enteredthe squish area 171 enters the swirl flow, and the remaining fuel thathas entered the region inside the cavity 31 enters the inside of theswirl flow. The fuel that has entered the swirl flow remains in theswirl flow in a period from the intake stroke to the compression stroke,and forms air-fuel mixture for CI combustion in the outercircumferential portion of the combustion chamber 17. The fuel that hasentered the inside of the swirl flow also remains inside the swirl flowin the period from the intake stroke to the compression stroke, andforms air-fuel mixture for SI combustion in the center portion of thecombustion chamber 17.

When the engine 1 operates in the high-load intermediate-rotation region(2), the excess air ratio K is preferably 1 or less in the air-fuelmixture in the center portion where the spark plug 25 is disposed, andthe excess air ratio K is 1 or less and preferably less than 1 in theair-fuel mixture in the outer circumferential portion. The air-fuelratio (A/F) of the air-fuel mixture in the center portion may be, forexample, in a range from 13 to the theoretical air-fuel ratio (14.7).The air-fuel ratio of the air-fuel mixture in the center portion may beleaner than the theoretical air-fuel ratio. Furthermore, the air-fuelratio of the air-fuel mixture in the outer circumferential portion maybe, for example, in a range from 11 to the theoretical air-fuel ratio,and may be preferably in a range from 11 to 12. When the excess airratio K in the outer circumferential portion of the combustion chamber17 is less than 1, an amount of fuel in the air-fuel mixture isincreased in the outer circumferential portion. Therefore, a temperaturecan be lowered by latent heat of vaporization of the fuel. The air-fuelratio of the air-fuel mixture in the entirety of the combustion chamber17 may be in a range from 12.5 to the theoretical air-fuel ratio, andmay be preferably in a range from 12.5 to 13.

In the succeeding injection 6032 that is performed at the terminal stageof the compression stroke, for example, fuel injection may be started at10° CA before the top dead center. By the succeeding injection beingperformed immediately before the top dead center, a temperature in thecombustion chamber can be lowered by latent heat of vaporization of thefuel. Low-temperature oxidation reaction of fuel injected in thepreceding injection 6031 proceeds in the compression stroke, and shiftsto high-temperature oxidation reaction before the top dead center.However, the succeeding injection 6032 is performed immediately beforethe top dead center, and a temperature in the combustion chamber is thuslowered, whereby shift from the low-temperature oxidation reaction tothe high-temperature oxidation reaction can be inhibited, and occurrenceof preignition can be inhibited. The ratio of the injection amount inthe preceding injection to the injection amount in the succeedinginjection may represent, for example, 95:5.

The spark plug 25 ignites the air-fuel mixture in the center portion ofthe combustion chamber 17 at or near the compression top dead center(reference numeral 6033). The spark plug 25 performs ignition, forexample, at or after the compression top dead center. Since the sparkplug 25 is disposed in the center portion of the combustion chamber 17,SI combustion of the air-fuel mixture in the center portion by flamepropagation starts by ignition by the spark plug 25.

In the high load region, an amount of injected fuel is increased, and atemperature in the combustion chamber 17 is also enhanced, so that theCI combustion is likely to start early. In other words, in the high loadregion, preignition of air-fuel mixture is likely to occur. However, asdescribed above, since a temperature in the outer circumferentialportion of the combustion chamber 17 is lowered by latent heat ofvaporization of fuel, start of the CI combustion immediately after sparkignition of air-fuel mixture can be prevented.

As described above, when the spark plug 25 ignites air-fuel mixture inthe center portion, a combustion speed is enhanced by high turbulentenergy to stabilize the SI combustion, and flame in the SI combustionpropagates in the circumferential direction by the strong swirl flow inthe combustion chamber 17. Thus, compression ignition of unburnedair-fuel mixture occurs at a predetermined position in thecircumferential direction in the outer circumferential portion of thecombustion chamber 17, to start the CI combustion.

According to the concept of the SPCCI combustion, air-fuel mixturestratifies in the combustion chamber 17, and strong swirl flow isgenerated in the combustion chamber 17, whereby the SI combustion can besufficiently performed before start of the CI combustion. As a result,generation of combustion noise can be inhibited, and, further, atemperature of the combustion does not become excessively high, andgeneration of NOx is also inhibited. Furthermore, variation in torquebetween cycles can be reduced.

Furthermore, since the temperature in the outer circumferential portionis low, the CI combustion is gentle, and generation of combustion noisecan be inhibited. Moreover, the CI combustion shortens the combustionperiod. Thus, torque is improved in the high load region and thermalefficiency is improved. Therefore, the engine 1 performs SPCCIcombustion in a region in which load is high, whereby fuel economyperformance can be improved while combustion noise is avoided.

When the engine 1 operates on the high rotation side in the high-loadintermediate-rotation region (2), the injector 6 starts fuel injectionin the intake stroke (reference numeral 6041).

In the preceding injection 6041 that starts in the intake stroke,similarly as described above, fuel injection may be started in theformer half period of the intake stroke. Specifically, in the precedinginjection 6041, fuel injection may be started at 280° CA before the topdead center. The preceding injection may end in the compression strokeafter the intake stroke. When injection in the preceding injection 6041starts in the former half period of the intake stroke, air-fuel mixturefor CI combustion can be formed in the outer circumferential portion ofthe combustion chamber 17, and air-fuel mixture for SI combustion can bealso formed in the center portion of the combustion chamber 17.Similarly as described above, the excess air ratio λ of the air-fuelmixture is preferably 1 or less in the center portion where the sparkplug 25 is disposed, and the excess air ratio λ of the air-fuel mixtureis 1 or less and preferably less than 1 in the outer circumferentialportion. The air-fuel ratio (A/F) of the air-fuel mixture in the centerportion may be, for example, in a range from 13 to the theoreticalair-fuel ratio (14.7). The air-fuel ratio of air-fuel mixture in thecenter portion may be leaner than the theoretical air-fuel ratio.Furthermore, the air-fuel ratio of the air-fuel mixture in the outercircumferential portion may be, for example, in a range from 11 to thetheoretical air-fuel ratio, and may be preferably in a range from 11 to12. The air-fuel ratio of air-fuel mixture in the entirety of thecombustion chamber 17 may be in a range from 12.5 to the theoreticalair-fuel ratio, and may be preferably in a range from 12.5 to 13.

When the number of revolutions of the engine 1 is increased, a time forreaction of fuel injected in the preceding injection 6041 is shortened.Therefore, the succeeding injection for inhibiting oxidation reaction ofthe air-fuel mixture can be omitted.

The spark plug 25 ignites the air-fuel mixture in the center portion ofthe combustion chamber 17 at or near the compression top dead center(reference numeral 6042). The spark plug 25 performs ignition, forexample, at or after the compression top dead center.

As described above, the air-fuel mixture stratifies, whereby combustionnoise is inhibited, and the SPCCI combustion can be also stabilized inthe high-load intermediate-rotation region (2).

(High-Load Low-Rotation Region (3))

When the engine 1 operates in the high-load low-rotation region (3), theengine 1 does not perform the SPCCI combustion but performs SIcombustion. The high-load low-rotation region (3) corresponds to thefirst high load region (C1) in the operation region map 701.

In FIG. 17, reference numeral 605 represents examples of a fuelinjection time (reference numeral 6051, 6052) and an ignition time(reference numeral 6053), and a combustion waveform (reference numeral6054) in the case of the engine 1 operating in an operation state ofreference numeral 605 in the high-load low-rotation region (3).

The EGR system 55 introduces EGR gas into the combustion chamber 17 whenan operation state of the engine 1 is in the high-load low-rotationregion (3). The engine 1 reduces an amount of EGR gas according to loadbeing increased. At full load, the EGR gas may be zero.

When the engine 1 operates in the high-load low-rotation region (3), theair-fuel ratio (A/F) of air-fuel mixture is the theoretical air-fuelratio (A/F≈14.7) in the entirety of the combustion chamber 17. The A/Fof the air-fuel mixture may be set so as to fall within a purificationwindow of the three-way catalyst. Therefore, the excess air ratio λ ofthe air-fuel mixture may be 1.0±0.2. When the air-fuel ratio of theair-fuel mixture is set to the theoretical air-fuel ratio, fuel economyperformance is improved in the high-load low-rotation region (3). Whenthe engine 1 operates in the high-load low-rotation region (3), theconcentration of fuel in the air-fuel mixture in the entirety of thecombustion chamber 17 may be set in a range from the excess air ratio λin the high-load intermediate-rotation region (2) to the excess airratio λ of 1, and preferably greater than the excess air ratio λ in thehigh-load intermediate-rotation region (2).

When the engine 1 operates in the high-load low-rotation region (3), theinjector 6 injects fuel into the combustion chamber 17 at timing in theintake stroke, and at timing in the retard period from the terminalstage of the compression stroke to the initial stage of the expansionstroke, in the operation region map 702 (reference numeral 6051, 6052).When fuel is injected in two parts, an amount of fuel to be injected inthe retard period can be reduced. By fuel being injected in the intakestroke (reference numeral 6051), a sufficient time in which air-fuelmixture is formed can be assured. Furthermore, by fuel being injected inthe retard period (reference numeral 6052), flow in the combustionchamber 17 can be enhanced immediately before ignition, and SIcombustion is advantageously stabilized.

The spark plug 25 ignites the air-fuel mixture at timing of or at timingnear the compression top dead center (reference numeral 6053) after fuelis injected. The spark plug 25 may ignite the air-fuel mixture, forexample, after the compression top dead center. SI combustion of theair-fuel mixture is performed in the expansion stroke. Since the SIcombustion starts in the expansion stroke, CI combustion does not start.

When the engine 1 operates in the high-load low-rotation region (3),swirl flow is weakened as compared to an operation in the high-loadintermediate-rotation region (2). In the operation in the high-loadlow-rotation region (3), the opening degree of the swirl control valve(SCV) 56 is greater than that in the operation in the high-loadintermediate-rotation region (2). The opening degree of the swirlcontrol valve 56 may be, for example, about 50% (that is, half-open).

In the upper view of FIG. 2, the axis of the hole of the injector 6 ispositioned so as to be shifted relative to the spark plug 25 in thecircumferential direction as indicated by an arrow drawn by an alternatelong and short dash line. Fuel injected from the hole flows in thecircumferential direction due to swirl flow in the combustion chamber17. The swirl flow allows the fuel to be immediately transported to aportion near the spark plug 25. The fuel can be vaporized while the fuelis transported to the portion near the spark plug 25.

Meanwhile, when swirl flow is excessively strong, fuel is caused to flowin the circumferential direction, and is moved away from the portionnear the spark plug 25, and the fuel cannot be immediately transportedto the portion near the spark plug 25. Therefore, when the engine 1operates in the high-load low-rotation region (3), swirl flow isweakened as compared to an operation in the high-loadintermediate-rotation region (2). Thus, since fuel can be immediatelytransported to the portion near the spark plug 25, ignitability ofair-fuel mixture can be improved and SI combustion can be stabilized.

(High Rotation Region (4))

When the number of revolutions of the engine 1 is great, a time forchanging the crank angle by 1° is shortened. Therefore, for example, inthe high rotation region in the high load region, as described above,stratifying of air-fuel mixture in the combustion chamber 17 becomesdifficult. When the number of revolutions of the engine 1 is great, theabove-described SPCCI combustion becomes difficult.

Therefore, when the engine 1 operates in the high rotation region (4),the engine 1 does not perform the SPCCI combustion but performs SIcombustion. The high rotation region (4) extends over the entire regionfrom the low load region to the high load region in the load direction.

In FIG. 17, reference numeral 606 represents examples of a fuelinjection time (reference numeral 6061) and an ignition time (referencenumeral 6062), and a combustion waveform (reference numeral 6063) in thecase of the engine 1 operating in the operation state of the referencenumeral 606 in the high rotation region (4).

The EGR system 55 introduces EGR gas into the combustion chamber 17 whenan operation state of the engine 1 is in the high rotation region (4).The engine 1 reduces an amount of EGR gas according to load beingincreased. At full load, the EGR gas may be zero.

When the engine 1 operates in the high rotation region (4), the swirlcontrol valve (SCV) 56 is fully opened. No swirl flow is generated inthe combustion chamber 17, and only tumble flow is generated. By theswirl control valve 56 being fully opened, filling efficiency can beenhanced in the high rotation region (4), and pump loss can be reduced.

When the engine 1 operates in the high rotation region (4), the air-fuelratio (A/F) of air-fuel mixture is basically the theoretical air-fuelratio (A/F=14.7) in the entirety of the combustion chamber 17. Theexcess air ratio λ of the air-fuel mixture may be 1.0±0.2. In the highload region, including the full load, in the high rotation region (4),the excess air ratio λ of the air-fuel mixture may be less than 1.

When the engine 1 operates in the high rotation region (4), the injector6 starts fuel injection in the intake stroke (see reference numeral6061). The injector 6 injects fuel at one time. By starting the fuelinjection in the intake stroke, air-fuel mixture can be homogeneously oralmost homogeneously formed in the combustion chamber 17. Furthermore,when the number of revolutions of the engine 1 is great, fuelvaporization time can be assured so as to be as long as possible, sothat unburned fuel loss can be reduced, and generation of soot can beinhibited.

After injection of fuel has been ended, the spark plug 25 ignites theair-fuel mixture at appropriate timing before the compression top deadcenter (see reference numeral 6062).

Other Embodiments

The technique disclosed here is not limited to application to the engine1 having the above-described configuration. Various configurations canbe adopted as the configuration of the engine 1.

For example, the cavity 31 may have a shallow bottom portion having thebottom shallower than the depressed portion 312, in a portion opposingthe spark plug 25. A part of fuel injected by the injector 6 is guidedby the shallow bottom portion to reach a portion near the spark plug 25.The fuel spray guided by the shallow bottom portion can reach the sparkplug 25 through a relatively short transport passage. In the first highload region (C1) of the operation region map 701 and the high-loadlow-rotation region (3) of the operation region map 702 as describedabove, fuel injected at the terminal stage of the compression stroke canbe immediately transported to a portion near the spark plug 25.

DESCRIPTION OF REFERENCE CHARACTERS

-   -   1 Engine    -   10 ECU (Controller)    -   17 Combustion Chamber    -   23 Intake Electric S-VT (State Quantity Setting Device, Variable        Valve Mechanism)    -   24 Exhaust Electric S-VT (State Quantity Setting Device,        Variable Valve Mechanism)    -   25 Spark Plug    -   49 Supercharging System (State Quantity Setting Device)    -   44 Supercharger    -   43 Throttle Valve (State Quantity Setting Device)    -   48 Air Bypass Valve (State Quantity Setting Device)    -   54 EGR Valve (State Quantity Setting Device)    -   55 EGR System (State Quantity Setting Device)    -   56 Swirl Control Valve (State Quantity Setting Device)    -   6 Injector

1. A control apparatus for an engine, comprising: an engine having acombustion chamber; an injector mounted to the engine and configured toinject fuel; a spark plug disposed so as to face an inside of thecombustion chamber; and a controller connected to the injector and thespark plug and configured to output control signals to the injector andthe spark plug, wherein the spark plug ignites air-fuel mixture atpredetermined ignition timing so that unburned air-fuel mixture combustsby autoignition after start of combustion of the air-fuel mixture by theignition, and the controller adjusts a heat amount ratio in accordancewith an operation state of the engine through change of the ignitiontiming, the heat amount ratio representing an index associated with aratio of an amount of heat generated when the air-fuel mixture combustsby flame propagation with respect to a total amount of heat generatedwhen the air-fuel mixture combusts in the combustion chamber.
 2. Thecontrol apparatus for the engine of claim 1, wherein the controlleroutputs a control signal to the spark plug so that, as a temperature inthe combustion chamber before start of compression decreases, theignition timing is advanced, thereby increasing the heat amount ratio.3. A control apparatus for an engine, comprising: an engine having acombustion chamber; a state quantity setting device provided to theengine and configured to adjust introduction of fresh air and burned gasinto the combustion chamber; an injector mounted to the engine andconfigured to perform injection; a spark plug disposed so as to face aninside of the combustion chamber; and a controller connected to thestate quantity setting device, the injector, and the spark plug, andconfigured to output control signals to the state quantity settingdevice, the injector, and the spark plug, wherein the spark plug ignitesair-fuel mixture at predetermined ignition timing so that unburnedair-fuel mixture combusts by autoignition after start of combustion ofthe air-fuel mixture by the ignition, the state quantity setting deviceadjusts operation quantity relevant to a temperature in the combustionchamber before start of compression, in response to a control signalfrom the controller, and the controller adjusts a heat amount ratio inaccordance with an operation state of the engine through change of theoperation quantity by the state quantity setting device and change ofthe ignition timing, the heat amount ratio representing an indexassociated with a ratio of an amount of heat generated when the air-fuelmixture combusts by flame propagation with respect to a total amount ofheat generated when the air-fuel mixture combusts in the combustionchamber.
 4. The control apparatus for the engine of claim 3, whereinwhen load on the engine is high, the controller outputs a control signalfor adjusting the operation quantity to the state quantity settingdevice so that the temperature in the combustion chamber before start ofcompression becomes lower than when the load is low, thereby increasingthe heat amount ratio.
 5. The control apparatus for the engine of claim4, wherein when the load on the engine is high, the controller outputs acontrol signal to the spark plug so that the ignition timing is advancedas compared to when the load is low, thereby increasing the heat amountratio.
 6. The control apparatus for the engine of claim 4, wherein thestate quantity setting device has a supercharging system provided to theengine and configured to perform supercharging with gas to be introducedinto the combustion chamber, and in response to a control signal fromthe controller, the supercharging system performs supercharging when theload on the engine is high, and does not perform supercharging when theload is low.
 7. The control apparatus for the engine of claim 6, whereinin a case of not performing supercharging using the superchargingsystem, the controller outputs a control signal for adjusting theoperation quantity to the state quantity setting device so that thetemperature in the combustion chamber before start of compressiondecreases as the load on the engine increases, thereby increasing theheat amount ratio.
 8. The control apparatus for the engine of claim 7,wherein in a case of not performing supercharging using thesupercharging system, the controller outputs a control signal to thespark plug so that the ignition timing is advanced as the load on theengine increases, thereby increasing the heat amount ratio.
 9. Thecontrol apparatus for the engine of claim 6, wherein in the case ofperforming supercharging using the supercharging system, the controlleroutputs a control signal to the spark plug so that the ignition timingis advanced as the load on the engine increases, thereby making the heatamount ratio constant with respect to change in the load on the engine.10. The control apparatus for the engine of claim 3, wherein thecontroller outputs control signals to the state quantity setting deviceand the injector, to set a G/F that represents an index associated witha weight ratio between total gas including burned gas in the combustionchamber, and fuel, such that the G/F is in a range from 18 to
 50. 11.The control apparatus for the engine of claim 10, wherein the controlleroutputs control signals to the state quantity setting device and theinjector, to set an excess air ratio λ of the air-fuel mixture to1.0±0.2.
 12. The control apparatus for the engine of claim 10, wherein astate inside the combustion chamber at the ignition timing satisfies atleast one of a condition that a temperature is in a range from 570 K to800 K, and a condition that a pressure is in a range from 400 kPa to 920kPa.
 13. The control apparatus for the engine of claim 10, wherein astate inside the combustion chamber at the ignition timing satisfies acondition that a swirl ratio is 4 or greater.
 14. The control apparatusfor the engine of claim 10, wherein a geometrical compression ratio ofthe engine is 13 or greater.